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STEAM ENGINES 



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PUBLISHERS OF BOOKS F O B^ 

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ENGINEERING EDUCATION SERIES 



STEAM ENGINES 



PREPARED IN THE 

EXTENSION DIVISION OF 
THE UNIVERSITY OF WISCONSIN 

, BY 

E. M. SHEALY 

ASSOCIATE PROFESSOR OF STEAM ENGINEERING 
THE UNIVERSITY OF WISCONSIN 



First Edition 



McGRAW-HILL BOOK COMPANY, Inc. 
239 WEST 39TH STREET. NEW YORK 



LONDON: HILL PUBLISHING CO., Ltd. 

6 «fe 8 BOUVERIE ST., E. C. 

1919 



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Copyright, 1919, by the 
McGraw-Hill Book Co., Inc. 




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TUB MAPLE PKH S S X O R K PA 



PREFACE 

This book on Steam Engines was written to be used as a text- 
book for correspondence students in the University of Wisconsin 
Extension Division. It is the third of a series of three textbooks 
designed for those students who are pursuing a general course in 
Steam Engineering, the other two being ''Steam Boilers" and 
''Heat." 

In this course in Steam Engines we aim to teach the fundamen- 
tal principles underlying the operation of the steam engine and 
to do this in as simple and nonmathematical a manner as possi- 
ble. This is particularly true with those parts which deal with 
thermodynamic principles. Enough of the practical features 
of steam engine operation has been given to illustrate the princi- 
ples, and it is hoped that operating engineers who take this 
course will be able to supplement from their own experience 
other applications of the principles presented. 

That part of the course dealing with valve gears has been made 
more complete than other sections because our experience shows 
that operating engineers usually do not understand the valve 
gear mechanism of their engines as well as they do other parts. 

Most of the material in the chapter on Lubrication was fur- 
nished by Mr. R. P. Tobin, Chief of the Technical Department 
of the Vacuum Oil Company and we take this opportunity to 
express our thanks for his aid. We wish to take this opportunity 
also to thank Mr. J. C. White, Chief Operating Engineer for the 
State of Wisconsin for very valuable suggestions as to the scope 
of the course and the outline to be followed, also for many useful 
hints and suggestions about writing the course, and for a care- 
ful and critical reading of the manuscript. 

E. M. SHEALY. 

Madison, Wis., 
November 12, 1918. 



CONTENTS 

CHAPTER I 

Principles of the Steam Engine 

Article Page 

Elementary Principles 1 

Parts of the Steam Engine 5 

Classification of Engines 6 

The Plain Shde Valve Engine 8 

Speed Regulation 11 

Automatic High Speed Engines 13 

CHAPTER II 

Corliss and other Engines 

Corhss Engines 17 

Nonreleasing Corliss Engine 23 

The Locomotive 25 

Marine Engines 26 

CHAPTER III 

Parts of the Steam Engine 

The Frame 27 

The Cylinder 30 

The Piston 37 

Stuffing Box 40 

The Crosshead 42 

Connecting Rods 45 

Crank and Crank Pin 47 

Bearings 48 

The Flywheel 51 

CHAPTER IV 

Heat, Work, and Pressure 

Force 53 

Work ■ 54 

Energy 54 

Heat 55 

Temperature 57 

Unit of Heat 57 

Mechanical Equivalent of Heat 58 

Specific Heat 58 

vii 



viii CONTENTS 

Article Page 

Power 58 

Atmospheric Pressure 59 

Vacuum 59 

Barometer 60 

Absolute and Gage Pressures 60 

Measuring Vacuum 61 

CHAPTER V 

Properties op Steam 

Formation of Steam 64 

Interpolation from Tables 68 

Wet Steam 69 

Superheated Steam 70 

CHAPTER VI 

Indicators 

Work Diagrams 79 

The Indicator 80 

Reducing Motions 87 

Indicator Diagrams 89 

Expansion of Steam 93 

Ratio of Expansion 94 

CHAPTER VII 
Indicated and Brake Horsepower 

Mean Effective Pressure 98 

Indicated Horsepower 102 

Engine Constant 103 

Brake Horsepower 103 

Mechanical Efficiency 106 

CHAPTER VIII 
Action of Steam in the Cylinder 

Cylinder Condensation 107 

The Uniflow Engine 112 

Measuring Cylinder Condensation 114 

CHAPTER IX 
Steam Engine Testing 

Principles 118 

Steam Consumption 119 

Steam Consumption from Diagram 119 

Duration of Engine Test 122 

Efficiency of Steam Engines 123 

Efficiency of a Perfect Engine 125 

Computations 126 

Calculating Results 127 

Duty of Pumps 131 



CONTENTS ix 



CHAPTER X 

The Slide Valve 

Article Page 

Steam and Exhaust Lap 133 

Valve Without Laps 134 

Valves With Lap 136 

Position of Crank and Eccentric 137 

Lead 138 

Angle of Advance 140 

Inside Admission Valve 141 

CHAPTER XI 

The Valve Diagram 

Valve Displacement 143 

Piston Position 144 

Position of Crank and Eccentric 145 

Valve Diagram 146 

CHAPTER XII 

Valve Setting 

General Considerations 157 

Placing an Engine on Center 159 

To Set Valves With Equal Leads 161 

Setting Valves for Equal Cut-off 162 

Types of SHde Valves 164 

CHAPTER XIII 

Shifting Eccentric and Meyer Valve 

Shifting Eccentric 171 

Effects Produced by Slide Valve 177 

Meyer Valve 179 

CHAPTER XIV 

Reversing Mechanisms 

Reversing Gears 183 

Stephenson Link Motion 184 

Walschaert Valve Gear 191 

Woolf Reversing Gear 196 

CHAPTER XV 

Corliss Valve Gears 

Advantages of the Corliss Valve 198 

Single Eccentric Valve Gear 200 

Setting Corhss Valves 202 



CONTENTS 



CHAPTER XVI 
Governing 

Article Page 

Governing 211 

Pendulum Governor 212 

Stability 214 

Shaft Governors ' 220 

Inertia Governor 222 

CHAPTER XVII 

Compound Engines 

Compounding 225 

Expansion of Steam 227 

Compound Engines 228 

Cross-compound Engines 230 

Tandem-compound Engines 232 

Cross-compound with Receiver 233 

Power of a Compound Engine 235 

Advantages and Disadvantages 239 

CHAPTER XVIII 

Condensing Apparatus 

Purpose of the Condenser 240 

Condensation of Steam 242 

Measuring Vacuum 243 

Forms of Condensing Apparatus 246 

Jet Condenser 247 

Siphon Condensers 248 

Barometric Condenser 249 

Surface Condensers 250 

High Vacuum Condensers 253 

Choice of a Condenser 253 

CHAPTER XIX 

Lubrication 

Friction 255 

Lubrication 255 

Principles of Lubrication 256 

Characteristics of Oil 259 

Testing Oils 260 

Gumming Test 260 

Flash and Fire Tests 260 

Acid Test 261 

Steam-engine Lubrication 261 



CONTENTS xi 

Article Page 

Lubricators 263 

Lubrication of Valves. Slide Valve 266 

Corliss Valves 267 

Piston Valves 267 

Poppet Valves 267 

Piston and Cylinders 268 

Piston and Valve Rods 268 

Influence of Operating Conditions 269 

CHAPTER XX 

Steam Turbines 
General Principles 271 



STEAM ENGINES 

CHAPTER I 
PRINCIPLES OF THE STEAM ENGINE 

Elementary Principles. — In the steam engine, heat energy is 
changed into mechanical energy. The pressure of steam is due 
to the heat which it contains. The steam pressure acts upon 
the engine piston, causing it to move, and thus changes the heat 
energy of the steam into mechanical energy. The kind of motion 
produced is a backward and forward motion of certain parts of 
the engine. This kind of motion is called a reciprocating motion. 
and the parts of the engine which have this kind of motion are 
called reciprocating parts. Other parts of the engine change the 
reciprocating motion into a rotary motion and thus the engine 
may turn a flywheel continuously and transmit the motion to 
other machines. 

Figure 1 is a drawing of a steam engine, simplified in order 
that its principles may be more readily understood. Practically 
all steam engines operate upon the same principles, hence this 
explanation will serve for all classes of engines. Only the frame, 
cylinder, piston, piston rod, crosshead, connecting rod, crank, 
shaft, and flywheel are shown here. The frame and cylinder are 
stationary parts and the others are moving parts. The piston, 
piston rod, and crosshead have a reciprocating motion, and the 
crank, shaft, and flywheel have a rotary motion. 

The piston is moved backward and forward by the pressure 
of the steam which is admitted first to one end of the cylinder and 
then to the other. The supply of steam is controlled by means of 
a valve operated by the engine so as to open and close at the 
proper time. This valve, which is an important part of the engine 
mechanism, is not shown in Fig. 1, but will be illustrated and 
described later. 

Referring to Fig. 1, imagine steam to be admitted to the right- 
hand end of the cylinder. The pressure of the steam acts upon 

1 



STEAM ENGINES 



the piston and moves it to the left. The distance which the 
piston travels from left to right or from right to left is called the 
stroke. The motion of the piston is transmitted through the 
piston rod to the crosshead, which has the same motion as the 
piston. One end of the connecting rod is connected to the cross- 
head and moves in a straight line with it. The other end of the 
connecting rod is connected to the crank pin, which moves in a 
circle about the center of the shaft; therefore, the connecting 
rod changes the straight line motion of the crosshead into the 
rotating motion of the shaft and flywheel. 



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Fig. 1. 

When the steam pressure forces the piston of Fig. 1 to the left, 
the crank pin rotates through the top half of its circular path, 
moving in the direction of the arrow. As the piston moves to the 
left, more steam enters the cylinder and maintains a constant 
pressure upon the piston. At a certain point in the stroke of the 
piston the valve closes and stops the supply of steam to the right- 
hand end of the cylinder. This point in the stroke is called the 
point of cut-off or simply cut-off. From the point of cut-off to the 
end of the stroke the steam behind the piston expands and its 
pressure diminishes. At the end of the stroke from right to left, 



PRINCIPLES OF THE STEAM ENGINE 3 

connection between the right-hand end of the cyhnder and the 
exhaust pipe is opened and the steam begins to be exhausted from 
the cyhnder. The point in the stroke at which steam begins to 
be exhausted from the cyhnder is called release. 

If the piston should stop just at the end of either stroke, the 
piston rod, connecting rod, and crank would be in a straight line, 
and the engine would be on center or on dead center, as it is some- 
times called. In this position steam pressure acting on the piston 
could not move it, since the pressure would be simply transferred 
to the bearings of the shaft and there would be no turning effect. 
However, after completing a stroke, the motion of the parts of an 
engine is sufficient to carry it past center and the steam pressure 
will then move the piston forward. 

At the beginning of the stroke from left to right, the valve 
admits steam to the left-hand end of the cylinder while keeping 
open the connection between the right-hand end and the exhaust 
pipe. The steam pressure now forces the piston towards the 
right, and the crank is forced through the bottom half of its 
circular path, still in the direction of the arrow, thus causing the 
shaft and flywheel to turn continuously in the same direction. 
As the piston moves towards the right, the low pressure steam in 
the right-hand end of the cylinder is forced out by the piston. 
Steam continues to be admitted to the left-hand end of the 
cylinder until the point of cut-off in this stroke, after which the 
steam is expanded behind the moving piston until the end of the 
stroke, when exhaust commences from the left-hand end. Just 
before the piston completes its stroke from left to right, the 
connection between the right-hand end of the cylinder and the 
exhaust pipe is closed and the steam then remaining in the right- 
hand end of the cylinder is compressed in order to furnish a 
cushion for the returning piston. The point in the stroke at 
which the exhaust passage is closed is called compression. 

When an engine passes through a regular series of operations 
and returns at regular intervals to its starting point, it is said to 
perform a cycle. The parts of the cycle are called events. The 
events in the cycle of a steam engine are admission, cut-off, 
release, and comjpression. The part of the cycle between the 
point of admission and the point of cut-off is called admission; 
the part between the point of cut-off and the point of release is 
called expansion; the part between the point of release and the 
point of compression is called exhaust; and the part between the 



4 STEAM ENGINES 

point of compression and the point of admission is called 
compression. 

The series of operations, admission, expansion, exhaust, and 
compression occurring in one end of a cylinder make up the cycle 
for that end. If the cycle is performed in only one end of the 
cylinder the engine is said to be single-acting, but if the cycle is 
performed in each end of the cylinder, as in the engine described 
above, the engine is said to be double-acting. Nearly all steam 
engines are double-acting, since a double-acting engine has about 
twice the power of a single-acting one of the same size. In a 
double-acting engine the cycles in both ends of the cylinder are 
being performed at the same time, admission and expansion of 
one cycle occurring in one end of the cylinder at the same time 
that exhaust and compression of the other cycle are occurring in 
the other end of the cylinder. 

The end of the cylinder which is towards the shaft or flywheel 
is called the crank end, and the opposite end, or the one furthest 
from the shaft or flywheel, is called the head end of the cylinder. 
The stroke of the piston from the head end of the cylinder to the 
crank end is called the forward stroke and the stroke from the 
crank end to the head end is called the return stroke. 

When the piston is at the end of its stroke it does not touch the 
head of the cylinder, a small amount of clearance between them 
being necessary. The space between the head of the cylinder 
and the piston (when it is at the end of its stroke), together with 
the volume of the ports, up to the face of the valves, is called 
the clearance volume, or simply the clearance. The clearance is 
expressed as a percentage of the volume displaced by the piston 
during a single stroke. For example, the clearance of an engine 
may be 12 per cent. If a 20" X 24'' engine is under consideration 
(meaning an engine having a cylinder 20 in. in diameter with a 
24 in. stroke) the area of its piston is 

.7854 X 202 = 314.16 sq. in. 
and the volume displaced during a single stroke is 
314.16 X 24 = 7539.84 cu. in. 

7539.84 . _. „^ 
or —yj2Q~ ^ 4.305 cu. ft. 

The clearance volume of this engine is 

12 per cent, of 4.305 or 
.12 X 4.305 = .5166 cu. ft. 



PRINCIPLES OF THE STEAM ENGINE 5 

Parts of the Steam Engine. — The different parts of a steam 
engine in their relation to each other are shown in Fig. 2, which 
represents a common for.m of steam engine. In this view the 




cyhnder is shown cut away in order to illustrate its interior 
construction. 

In Fig. 2, 1 is the foundation of the engine; 2 is the frame; 



/ 



6 STEAM ENGINES 

3 is the cylinder; 4 is the head end cyHnder head; 5 is the crank 
end cyHnder head; 6 is the piston; 30 is the piston rod; 8 is the 
crosshead; 9 and 9 are the crosshead guides; 10 is the connecting 
rod; 12 and 12 are the cranks, this being a center crank engine; 
31 is the shaft; 13 is the flywheel; 14 is the eccentric; 15 is the 
eccentric strap; 16 is the eccentric rod; 17 is the valve stem guide; 
18 is the valve stem; 19 is the valve which, in this case, is a slide 
valve; 20 and 21 are the steam ports; 22 is the exhaust port; 
23 is the steam chest, which is connected to the steam supply 
pipe; 24 is the exhaust pipe, which is connected to the exhaust 
port; 25 is the piston rod stuffing box; 26 is the valve stem stuffing 
box; 27 is the steam chest cover plate; and 28 is the lagging or 
covering for the cylinder. 

Classification of Engines. — Steam engines may be divided into 
three classes depending upon their type of valve or method of 
controlling the speed, and these classes include practically all 
kinds of engines. These three classes are: 
The plain slide valve engine 
The automatic high speed engine 
The Corliss engine 
Any of the above types may, be classified in several other ways, 
among which are: 

According to the position of the cylinder, as horizontal and 
vertical; 

According to the number of cylinders in which the steam is 
expanded, as simple, compound, triple expansion, and quadruple 
expansion; 

According to the manner of handling the exhaust steam as, 
condensing and noncondensing. 

A horizontal engine is one whose cylinder is placed in a hori- 
zontal position as illustrated in Fig. 2, while a vertical engine 
is one having the cylinder placed vertically and directly above the 
shaft as shown in Fig. 3. Some of the largest engines are a 
combination of horizontal and vertical, having one horizontal 
and one vertical cylinder, and the connecting rods from each 
of these connected to a single crank. Phis arrangement is used 
in order to develop a large amount of power in a small space. 
A simple engine is one in which the steam is expanded in only 
one cylinder. In a compound engine the steam is first expanded 
in one cylinder and the exhaust from this cylinder is led to a 
second cylinder where it is expanded further. In a triple expan- 



PRINCIPLES OF THE STEAM ENGINE 7 

sion engine the total expansion of the steam is divided into three 
parts, each being performed in a separate cyUnder, while in a 
quadruple expansion engine the total expansion of the steam is 
divided into four parts, each being performed in a separate cylin- 
der. The general names of multiple expansion or compound are 
used to designate any engine in which the expansion is performed 
in more than one cylinder. The reasons for dividing the expan- 




FiG. 3. 

sion of the steam into parts and also the construction of multiple 
expansion engines will be taken up in a later chapter. 

A condensing engine is one in which the exhaust steam is 
changed into water. Since the water thus formed occupies 
less space than the exhaust steam the back pressure against which 
the piston must make its return stroke is reduced. In a noncon- 
densing engine the exhaust steam is turned into the atmosphere 
and the piston must return against the pressure of the atmosphere 
plus enough pressure to force the exhaust steam through the 
exhaust pipe and ports. This pressure may amount to from 



8 



STEAM ENGINES 



18 to 20 lbs. per sq. in. absolute, or 3 to 5 lbs. per sq. in. above 
atmospheric pressure. 

The Plain Slide Valve Engine. — This type of engine is named 
from the kind of valve which is used to distribute steam to the 




o 



two ends of its cylinder, this type of valve being called sl slide valve. 
Figure 2 illustrates a common form of plain slide valve engine. 
The slide valve mechanism is also illustrated in Fig. 2 where it is 
shown as a part of the engine. The valve and mechanism is 
again shown in Fig. 4, but in this case the parts are placed differ- 



PRINCIPLES OF THE STEAM ENGINE 



9 



ently than in Fig. 2 in order to better show the operation of the 
mechanism. In Fig. 4 the valve is marked 1; 2 is called the 
valve seat; 3 is the face of the valve; 4 is the valve rod; 5 is the 
eccentric rod; 6 is the eccentric; 7 is the shaft; and 8 is the 
eccentric strap. 

The slide valve moves backward and forward over both steam 
ports and the exhaust port, and is thus able to control the supply 
of steam to both ends of the cylinder and also the exhaust from 
both ends. The valve is given its backward and forward motion 
by the eccentric which is fastened by set screws to the shaft. 
The eccentric consists of a circular disk of iron having its center, 
O, in Fig. 4, at some distance from the center of the shaft N. The 
center of the eccentric thus moves in a circle about the center 





Fig. 5. 



of the shaft and in this way has a motion similar to that of a 
crank with a length NO. A crank which would give the same 
motion as the eccentric is shown by the dotted lines. The motion 
of the eccentric is transmitted to the eccentric rod by means of 
a strap, 8, which passes around the eccentric. The eccentric 
rod corresponds to a connecting rod and it changes the circular 
motion of the eccentric into a reciprocating motion, which is 
transmitted to the valve by the valve rod. 

The action of the slide valve can best be explained by consider- 
ing the series of operations which occur in one end of the cylinder, 
remembering that similar operations are occurring in the other 
end but at a different time. In Fig. 4 the piston is at the head 
end of the cylinder and just beginning its forward stroke, the 
shaft turning in the direction of the arrow. 

In the position shown in Fig. 4, the valve is opening to admit 



10 



STEAM ENGINES 



steam to the head end of the cyhnder. The steam pressure moves 
the piston towards the right and the valve is opened wider, which 
allows steam to flow into the cyhnder more freely. The valve 
soon reaches the end of its travel towards the right and begins 
to move towards the left, closing the head end steam port. Fig- 
ure 5 shows the valve just as it closes the head end steam port 
which cuts off the supply of steam to the cylinder. It will be 
seen from this figure that the piston has still some distance to 
go before completing its forward stroke. Figure 5 shows the 
position of the valve and piston at cut-off. The valve continues 
to move towards the left, keeping the steam port closed, and the 
steam expands behind the piston, pushing it towards the right. 
By the time the piston reaches the end of its forward stroke the 
inner edge of the valve begins to uncover the head end steam 





Fig. 6. 

port, as shown in Fig. 6, and gives the event known as release. 
This opens communication between the head end of the cyhnder 
and the exhaust port; then if the steam still has any pressure 
above that of the atmosphere, this pressure immediately drops 
to the exhaust pressure. 

Steam is now admitted to the crank end of the cylinder and the 
piston moves towards the left, pushing the spent steam in the 
head end of the cyhnder into the exhaust pipe. This part of the 
return stroke gives exhaust from the head end of the cylinder. 
The valve continues to move towards the left, opening the exhaust 
wider and wider, until it reaches the end of its travel, when it 
begins to move towards the right and to close the exhaust port. 
When the piston has reached the point in its return stroke shown 
in Fig. 7, the valve has moved far enough to the right to close the 



PRINCIPLES OF THE STEAM ENGINE 



11 



exhaust port. From this point to the end of the return stroke, 
the exhaust passage remains closed, and the piston compresses 
the steam which remains in the head end of the cyhnder so that 
at the end of the stroke the clearance volume of the cylinder is 
filled with high pressure steam. This completes the cycle in 
the head end of the cylinder. 

By referring to Figs. 4, 5, 6, and 7 it will be seen that the 
valve is so constructed that at the same time admission and 
expansion are occurring in the head end of the cylinder, exhaust 
and compression are occurring in the crank end; and at the same 
time that exhaust and compression are occurring in the head end, 
admission and expansion are occurring in the crank end. Thus 
the cycles for both ends of the cylinder are performed in their 





Fig. 7. 



proper order and the engine made to run continuously by means 
of a single slide valve and a single eccentric. 

Speed Regulation. — The speed of a steam engine may vary in 
two ways and from two different causes. First, there may be a 
variation during each stroke on account of the changing steam 
pressure in the cylinder. In the first part of the stroke when the 
steam pressure is high there is a tendency for the speed to increase 
and in the last part of the stroke when the steam pressure has 
been reduced by expansion there is a tendency for the speed to be 
lower. This variation of speed is entirely independent of the load 
carried by the engine and it would occur whether the load upon 
the engine was large or small. 

The variation of speed during a single stroke is counteracted 
by the action of the flywheel. It is a well-known fact that when a 



12 STEAM ENGINES 

heavy object is moving it is difficult to change its speed. This 
property of a body is called its inertia. The inertia of the flywheel 
is used to counteract the changes of speed during a single stroke. 
The flywheel absorbs energy during the first part of the stroke 
when the speed tends to increase and gives it out again during 
the last part of the stroke when the speed tends to decrease. 
A heavy flywheel will keep the speed of an engine steadier than a 
light one and weight concentrated at the rim is more effective 
than weight nearer the hub, hence flywheels which are intended 
to steady the speed are usually made with a very heavy rim. 

Besides the variation of speed mentioned above, there is also 
a variation due to changes in the load which the engine carries. 
Changes in the load affect the speed for a longer period than a 
single stroke and they cannot be controlled by the action of the 
flywheel. If the amount of steam supplied to the cylinder at 
each stroke is constant and the load increases, the speed of the 
engine will decrease until the power developed in the cylinder 
balances the load on the engine, and if the load decreases the 
speed increases until the power developed in the cylinder again 
balances the load. The load on most engines is varying all the 
time and, as it is desirable to keep the speed constant, some means 
must be provided for varying the amount of steam supplied to 
the engine according to the load it is carrying, so that when the 
load increases the amount of steam supplied will be greater and 
when the load decreases the amount of steam supplied will be 
less. This is called governing the engine, and the mechanism for 
controlling the steam supply is called a governor. 

The governor of a steam engine may control the speed by 
changing the steam pressure or by changing the volume of steam 
admitted to the cylinder during the period of admission. In the 
first method the governor operates a throttle valve placed in the 
main steam pipe where it enters the steam chest and the partial 
closing of this valve reduces the steam pressure. By this means 
the governor controls the steam pressure acting upon the piston 
to agree with the load on the engine. This kind of governor is 
called a throttling governor. 

In the second method mentioned above the governor is 
arranged so as to change the point of cut-off and thus change the 
volume of steam admitted to the cylinder to agree with the load. 
When the load increases the governor makes the point of cut-off 
occur later in the stroke thus admitting more steam to the 



PRINCIPLES OF THE STEAM ENGINE 13 

cylinder, and when the load decreases the point of cut-off occurs 
earlier, admitting a smaller volume of steam to the cylinder. 

An engine whose speed is regulated by throttling or reducing 
the pressure of the steam supply uses a large amount of steam in 
proportion to the work it performs, or is inefficient, because the 
full pressure of the steam is used only when the load is greatest 
and for any smaller load a portion of the steam pressure is wasted. 
The method of governing in which the volume of steam admitted 
to the cylinder is changed is more economical because all of the 
steam admitted to the cylinder is used at the full boiler pressure. 

In the plain slide valve engine the position of the eccentric 
determines the part of the stroke at which cut-off occurs. Since 
the eccentric is fastened to the shaft the point of cut-off occurs at 
a fixed point in the stroke, therefore the speed is governed by the 
throttling method. 

The plain slide valve engine is usually designed to run at 
slow or medium speeds, with a stroke somewhat greater than the 
diameter of the piston. Most of them are of small size, since 
they are uneconomical in the use of steam. Phey are simple in 
construction and cheap in cost, hence are much used where only 
a small amount of power is needed and where expert attendance 
is not obtainable. With ordinary care, they last a long time and 
do not easily get out of order. This type of engine uses from 35 
to 60 pounds of steam per hour for each horsepower developed. 

Automatic High Speed Engines. — The type of engine known 
as the automatic high speed engine is also a slide valve engine, 
but its valve differs somewhat from that used in the plain slide 
valve engine and it also differs in other details of construction. 

The valves used on the automatic high speed type of engine are 
of much better construction than those used on the plain slide 
valve engine. One kind of valve commonly used on these engines 
is illustrated in Fig. 8. This kind of valve is known as a balanced 
valve because it has a balance plate which prevents the steam 
pressure from acting on the back of the valve. The plain slide 
valve, which has no balance plate, has the steam pressure in the 
steam chest acting on the entire area of the back of the valve. 
As the area of the valve is large, the valve is pressed against its 
seat with an enormous pressure, requiring a large amount of 
work in moving the valve. The large area of the valve also 
makes it difficult for oil to get between the valve and its seat to 
lubricate it. Some valves have balance platis which cover about 



14 



STEAM ENGINES 



80 per cent, of the area of the valve, leaving 20 per cent, of its 
area upon which the steam pressure may act. This part of the 
steam pressure is sufficient to keep the valve properly seated 




Fig. 8. 

and is not enough to cause undue friction. The balance plate 
is adjustable to allow for wear of the valve, being held in position 
by screws which press against the sides of the steam chest and 
hold it in place. 

f^///////y/^ /yyy/y^-v^ 




Fig. 9. 



Another kind of valve, commonly used on automatic high 
speed engines, is illustrated in Fig. 9. This form of valve is 
made in the shape of a spool and is like a slide valve which has 
been curved into a cylindrical form. Its motion is the same as 
that of the plain slide valve. In the valve shown here, which is 



PRINCIPLES OF THE STEAM ENGINE 



15 



called a piston valve, steam enters the steam chest at the central 
part of the valve and is admitted to the cylinder past the inner 
edges of the valve, exhaust taking place past the outer edges. 
This is the reverse of the manner in which a plain slide valve 
admits and exhausts steam. The steam ports completely sur- 
round the valve so that a large port opening is secured with a 
small movement of the valve. This valve does not require a 
balancing plate because the steam pressure acts on all sides 
of the valve equally, thus making it perfectly balanced. 

In the plain slide valve engine the volume of steam admitted 
to the cylinder at each stroke of the piston is constant, the speed 
of the engine being controlled by changing the admission pressure 




Fig. 10. 

of the steam. In the automatic high speed engine, the admission 
pressure of the steam is constant, the speed being controlled by 
changing the point of cut-off, and thus regulating the volume of 
steam admitted to the cylinder to suit the amount of work being 
done. The latter is the more economical method of controlhng 
the speed because the full steam pressure is utilized all the time. 
The point of cut-off is changed in the high speed engine by means 
of a governor which is attached to the eccentric in such manner 
that the eccentric may be shifted around on the shaft. The 
mechanism for doing this will be fully described in the chapter 
on governors. 

Figure 10 is a side view of an automatic high speed engine, and 
serves to show the proportion of its parts. It will be noticed 
that the engine is self-contained, that is, it rests on a bedplate 



16 STEAM ENGINES 

which forms a foundation for it. These engines are usually 
short, the parts being grouped closely. As compared with a 
plain slide valve engine, the automatic high speed engine has a 
shorter cylinder and connecting rod in proportion to the diameter 
of the cylinder. It is made in these proportions because a piston 
speed of about 500 to 700 feet per minute is desirable for all 
classes of engines and in order to secure this piston speed the 
length of stroke must be short if the number of revolutions per 
minute is large. 

The automatic high speed engine is made for speeds up to 
about 350 revolutions per minute and in size up to about 600 
horsepower. It has a close speed regulation at all loads and is, 
therefore, well adapted for direct connection to electric genera- 
tors, a class of work which requires high speeds and 
close speed regulation. These engines are also often con- 
nected to line shafting by means of belts, and used for general 
power purposes. They are more efficient than the plain slide 
valve engine, using from 30 to 40 pounds of steam per hour per 
horsepower. 



CHAPTER II 



CORLISS AND OTHER ENGINES 



Corliss Engines. — The Corliss engine is an entirely different 
type from either the plain slide valve or the automatic high speed 
type. This type of engine, like the others, is named from its 
type of valve, which is known as the Corliss valve. 

The Corhss valve is cylindrical in shape and is placed across 




Fig. 11. 



the cylinder instead of parallel with it. There are four of these 
valves for each cylinder, one admission valve for each end of the 
cylinder and one exhaust valve for each end. Each of these 
valves rotates about its axis instead of moving backward and 




VALVE STEM 



WORKING EDOE 

Fig. 12. 



forward parallel with its axis, as does the piston tj^pe of slide 
valve. The Corliss valve does not turn through a complete 
revolution, but oscillates back and forth through an angle only 
large enough to uncover the port. 

A cross section of a cylinder fitted with Corliss valves is shown 

17 



18 



STEAM ENGINES 



in Fig. 11, and a view of one of the valves removed from the 
cyUnder is shown in Fig. 12. It will be observed that, by this 
arrangement of valves, the ports are made short and the clear- 
ance reduced. The ports extend across the cylinder and are 
about equal in length to the diameter of the cylinder. Steam 
pressure acts upon the backs of the valves keeping them pressed 
against their seats, but the friction is small since the valves are 
small and they travel only a short distance in rotating through 
a small angle. The travel of the valve, in order to gain a full 
port opening, is sometimes further decreased by having two open- 
ings through the valve, or making them ''double ported" instead 
of having the steam pass only one edge of the valve. By having 




Fig. 13. 

two ports through each valve, the valve need travel only one- 
half as far for a given port opening as would a single ported 
valve. 

The Corhss valve is operated from an eccentric, the same as 
a sUde valve, but it differs from the sHde valve in that the admis- 
sion valves are connected to the eccentric only at those times 
when they are being opened. The exhaust valves are connected 
to the eccentric at all times, but the mechanism for moving them 
is such that they move through a very small angle, thus reducing 
friction. 

The mechanism by which the Corliss valve is operated is shown 
in Fig. 13 in which some of the rods are represented by single 
lines in order to make the drawing clearer. As shown in this 



CORLISS AND OTHER ENGINES 



19 



drawing, the admission valves are at the top of the cyHnder and 
the exhaust valves at the bottom. 

The eccentric rod is connected at F to the rocker arm B, which 
is pivoted at D, so that the end, G, of the rocker arm swings back 
and forth through a small angle. A wheel, 4, called a wrist 
plate, which is free to turn, is placed on the side of the cylinder, 
and a point, H, on its rim is- connected by a rod to the upper end, 
G, of the rocker arm. The wrist plate, therefore, oscillates 
back and forth through the same angle as the upper end of the 
rocker arm. All of the valves are connected to the wrist plate 
by means of the rods 5, 5, and 7, 7, and take their motion from it. 

The exhaust valves are fitted with round stems which pass 
through stuffing boxes in the side of the cylinder, so they are 




Fig. 14. 

free to turn without allowing steam to leak past them. To the 
ends of the exhaust valve stems are fitted cranks, 8 and 8, which 
are connected by rods, 7 and 7, to the rim of the wrist plate. 
The oscillation of the wrist plate causes these cranks to move 
back and forth through a small angle, thus opening and closing 
the exhaust valves. 

The admission valves are not connected directly to the wrist 
plate, but are arranged so that they are connected to it while the 
v^alves are being opened, and disconnected while the valves are 
being closed. The device by which this is accomphshed is shown 
in detail in Fig. 14. In this figure, the round valve stem project- 
ing through its stuffing box is shown at 9 in the figure at the left. 
The end of the valve stem is supported in a yoke, B, called a 
bonnet, which is bolted to the side of the cyHnder. The end of 
the bonnet is finished round and upon it are fitted two cranks, 



20 STEAM ENGINES 

6 and 11, which are free to turn. The crank 6 is L-shaped and is 
called a bell crank. One arm of the bell crank is connected by a 
rod, 5, to the wrist plate; the other end carries a V-shaped hook, 
10, called a steam-hook, which is pivoted at J. The end of the 
valve stem carries a crank, 9, which is keyed to it beyond the end 
of the bonnet. The end of this crank has upon it a small square 
steel block, P, and there is another small steel block, L, upon the 
end of the hook-claw. As the wrist plate oscillates back and 
forth the bell crank, 6, moves over far enough so its steel block, L, 
hooks under the steel block, P, on the valve stem crank. As 
the bell crank moves back, the valve stem crank is pulled with it, 
thus opening the valve. At a certain point in the backward 
movement of the bell crank the governor causes the steam-hook, 
10, to release the valve stem crank and the valve is closed, which 
causes steam to be cut off from that end of the cylinder. The 
crank, 11, is carried by the bonnet and is free to turn upon it. 
This crank is connected to the governor by the rod 12. The 
crank 11 has a small steel block, Ci, fastened on its hub, and the 
position of this steel block is determined by the position of 
the governor which is connected to the crank 11. As the V- 
shap'ed hook-claw moves backward after picking up the valve 
stem crank, the end, T, of the hook-claw strikes the steel block 
on crank 11 and causes the hook-claw to release the valve stem 
crank. Referring to Fig. 13, it will be seen that the wrist plate 
is near the extreme of its travel towards the right and that the 
hook-claw on the right-hand end is just ready to pick up the 
valve stem crank. At the same time the valve on the left-hand 
end has been opened and is ready to be released by the governor. 

The admission valves are made to close quickly by means of a 
dashpot, which is shown at 15 in Fig. 13. The dashpot consists 
of a cylinder 16 and a piston P^. The piston is connected by the 
rod 14 to the valve stem crank so that as this crank is raised upon 
opening the valve, the piston in the dashpot is also raised. Rais- 
ing the dashpot piston causes a vacuum in the dashpot which 
exerts a powerful suction on the piston -and closes the valve 
quickly when it is released, thus cutting off the supply of steam 
suddenly. 

It will be seen from the above description of the valve mechan- 
ism that the method of governing the speed of the Corliss engine 
is by changing the volume of steam admitted to the cylinder at 
each stroke which is the most economical method of governing. 



CORLISS AND OTHER ENGINES 21 

The automatic high speed engine uses this method of governing, 
also; but in this case the arrangement of the valve mechanism is 
such that when the point of cut-off changes, the point of com- 
pression changes also, as will be seen after this type of valve 
mechanism is studied in a later chapter. In the automatic 
engine the valve closes somewhat gradually which causes the 
admission pressure to decrease gradually up to the point of 
cut-off, while in the Corliss engine the valve opens wide at 
admission, remains wide open during admission, and closes 
suddenly at cut-off, thus giving full steam pressure upon the 
piston during the entire admission. At light loads the valve 
does not open full port area, but it generally opens enough to 
give full initial pressure. Also, since the steam valves are inde- 
pendent of the exhaust valves, the point of compression does not 
change when the point of cut-off changes in regulating the speed. 
On some Corliss engines the exhaust valves are not only inde- 
pendent of the steam valves but they are operated from a separate 
eccentric. The reasons for this will be explained fully in a 
later chapter. 

There are many modifications of the Corliss valve mechanism, 
but the principle is the same in all of them, that is, the valve is 
opened quickly by the eccentric, is disengaged from the eccentric 
at the proper time by the governor, and closed quickly by a 
vacuum dashpot. 

The valve mechanism is the distinctive feature of the Corliss 
engine, as aside from it this engine differs but little from any 
other type of slow speed engine. The invention of the Corliss 
valve mechanism is the greatest development that has taken place 
since the invention of the present form of steam engine by James 
Watt in 1769. The high efficiency of the Corliss engine is due 
to its form of valve mechanism more than to anything else as it 
permits an early cut-off with large expansion of the steam and 
does not throttle the steam pressure during admission or exhaust. 

A general view of a Corliss engine is shown in Fig. 15, which 
serves to give an idea of the proportion of its parts. It will be 
observed that this type of engine has a somewhat longer cylinder 
in proportion to its diameter and also a longer connecting rod 
than other types of engines. This gives the whole engine an 
appearance of considerable length in proportion to its height. 

The complicated valve mechanism used on the Corliss engines 
makes it necessary to run them at a relatively low speed in order 



22 



STEAM ENGINES 







CORLISS AND OTHER ENGINES 23 

for the various parts to adjust themselves and for the dash pot 
piston to work properly. In order to maintain a proper piston 
speed with the slow rate of revolution, the cylinder is made long 
in proportion to its diameter. Corliss engines are rarely made to 
run at a higher speed than 100 to 125 revolutions per minute, and 
in the larger sizes the speed is even slower than this. 

The Corliss engine is the most efficient type of steam engine. 
It rarely uses over 25 pounds of steam per hour per horsepower 
and in the larger sizes its steam consumption is much less than 
this. This type of engine is made in sizes from 100 to 12,000 
horse power. The Corliss engine is particularly adapted to 
running mills and for other power purposes on account of its 
smooth running qualities and its close speed regulation. The 
speed of the Corliss engine is controlled by changing the point 
of cut-off to suit the load, thus controlling the volume of steam 
admitted to the cylinder, the admission pressure of the steam 
remaining constant. 

Besides the three main types of engines mentioned above there 
are various modifications of these types which are used extensively 
and which are important on this account. Some of the principal 
ones of these are described below. 

Nonreleasing Corliss Engine. — The nonreleasing Corliss engine 
is a type of medium and high speed engine which has been devel- 
oped in recent years. It is a combination of the automatic high 
speed and the Corliss engines and it is used in the same kind of 
service as the automatic engine, that is, for direct connection of 
electric generators and for general power purposes by belting to 
line shafting. On account of its high speed this engine has the 
general shape and proportions of the automatic high speed engine, 
as will be seen from Fig. 16, which shows one of these engines. 

Any engine which runs at a high speed must have a positive 
connection between the eccentric and valve because there would 
not be time for a disengaging mechanism to operate properly. 
For this reason the nonreleasing Corliss engine has its admission 
valves, which are of the Corliss type, connected directly to a 
reach rod which is operated by an eccentric rod from the eccen- 
tric connected to the governor. With the Corliss type of valve 
connected directly to the governor eccentric, the nonreleasing 
Corliss engine has some of the characteristics of the Corliss 
engine and some of the automatic high speed engine. 

The method of governing the speed of the nonreleasing Corliss 



24 



STEAM ENGINES 



engine is the same as that used with the automatic high speed 
engine, that is, the point of cut-off is changed by shifting the 
eccentric which operates the admission valves around on the 
shaft. This does not change the point of compression because 
the exhaust valves are operated by a separate eccentric which is 
fastened to the shaft. This gives a steam distribution to the 
cylinder even better than that given by the automatic high speed 
engine with its slide valve, because release and compression 
can be fixed at the most advantageous points and they remain 
unchanged when the cut-off is changed. A double ported valve, 
which reduces the necessary movement of the valve for the same 
port opening and also further reduces the friction, is used in some 
of these engines. The type of valve used on this engine permits 




Fig. 16. 



short ports which reduce the clearance volume and thereby 
increase the economy of the engine. However, at least 6 per cent, 
clearance is needed for quiet running. 

In a sHde valve engine the same valve is used both for admission 
and exhaust. The exhaust steam, having a lower temperature 
than the admission steam, chills the valve so that when the next 
admission occurs, a part of the admission steam is condensed by 
coming in contact with the cooler valve. This condensation, 
though small is avoided in the nonreleasing Corliss engines since 
there are separate admission and exhaust valves. 

Since the nonreleasing Corhss engine combines some of the 
advantages of the Corliss engine with those of the automatic 
high speed engine, its steam consumption per horsepower per 
hour is slightly greater than that of the Corliss engine but less 



CORLISS AND OTHER ENGINES 25 

than that of the automatic high speed engine. It is made in 
sizes up to about 600 horsepower and for speeds up to about 
350 revolutions per minute. 

The Locomotive. — ^The locomotive, familiar to everyone, is a 
type of plain slide valve engine, differing from the ordinary plain 
slide valve engine in having the valve mechanism arranged so 
that the direction of rotation may be reversed at will. This type 
of valve mechanism will be studied in detail in a later chapter. 
Instead of having a single engine the locomotive has two complete 
engines, one on each side connected to the main driving shaft 
by cranks placed 90° apart so the locomotive may be started 
from any position even if one engine is on dead center. 

The valves used on locomotive engines are of the balanced 
type, some of them being flat valves with balance rings or plates 
on the back, and some being cylindrical or piston valves which are 
completely balanced. In both cases the usual location of the 
valve is on top of the cylinder where it is accessible to the 
engineer. 

Unlike the engines perviously mentioned, the locomotive has 
no flywheel, its function being performed by the driving wheels 
and by the weight of the boiler resting on them. Neither has the 
locomotive engine a governor since it is not intended to run at 
constant speed. Its speed is controlled by hand to suit the load 
which the engine is pulling. There are two means of regulating 
the speed : first by means of a hand-operated throttle valve which 
controls the pressure of the steam admitted to the cylinders; 
and second, by changing the point of cut-off, which may be done 
by means of the reversing mechanism. 

Until recent years the locomotive consisted of two simple 
engines. The demand for greater power has brought about the 
development of the compound locomotive in which the steam is 
first expanded in a high pressure cylinder and then in a low 
pressure cylinder, just as in the stationary compound engine. 
In some locomotives the high pressure cylinder is on one side and 
the low pressure cylinder on the other. This arrangement re- 
quires some provision whereby high pressure steam may be ad- 
mitted to the low pressure cylinder in order to start when the high 
pressure side is on center. Another common arrangement of 
the cylinders is to place oiie high and one low pressure cylinder 
on each side of the locomotive, thus making the locomotive 
consist of two complete compound engines. In this arrangement 



26 STEAM ENGINES 

of cylinders, the high pressure cylinder is placed directly above 
and parallel to its low pressure cylinder, the piston rods from both 
cylinders connecting to a single crosshead. This gives a more 
compact and powerful engine than the other arrangement of 
compound cylinders. 

A locomotive is a complete power plant in itself consisting of 
both boiler and engine, and, in some cases, of a feed water heater 
and superheater also. The amount of power developed is large 
compared with the size of the boiler and engine, being as much as 
2000 Hp. in some cases. When it is considered that this power 
is sometimes developed with a steam consumption of about 20-24 
pounds of steam per horsepower per hour, the efficiency of these 
machines is wonderful. 

Marine Engines. — Engines used on steamships form another 
distinct class of steam engines. These are also of the plain slide 
valve type and, like the locomotive, are provided with a mechan- 
ism for reversing the direction of rotation of the engine. Marine 
engines are vertical and are invariably multiple expansion in 
order to secure a large amount of power within a small space. 
The use of several cranks also gives a more uniform turning effort 
to the shaft. The engines are compound, triple or quadruple 
expansion depending largely upon the size of the engine. The 
cylinders are placed directly above the crank shaft with their 
axes vertical. In the larger vessels two propellers are used, 
each one on a separate shaft driven by a separate engine. 

There is no need for a governor on a marine engine as the resist- 
ance offered by the water to the revolving propeller increases as 
the speed increases and this prevents the engine from racing. 
Slower speeds are secured by partly closing a throttle valve which 
reduces the pressure of the steam supplied to the cylinders. 



CHAPTER III 
PARTS OF THE STEAM ENGINE 

The Frame. — The frame of an engine supports all of the work- 
ing parts and holds them in their proper relative positions. 
The form of the frame, especially on the larger sizes of engines, 
is determined by the type of engine and the purpose for which it is 
to be used; thus a Corliss engine used on rolling mill work would 
have an entirely different kind of frame than would the same type 
of engine when used for general power purposes, such as supplying 
power for a factory. 

The frame of the automatic high speed engine is the simplest 
of all engine frames. This frame, as shown in Figs. 10 and 16, 
is made to rest on a cast-iron sub-base instead of on a masonry 
foundation, hence the bottom edge of the frame is made in the 
form of a rectangle to fit the base. As most of these engines are 
of the center crank type, the whole frame is of rectangular shape, 
but is smaller at the top than at the bottom, as this shape is best 
adapted to supporting the bearings at each side, and to leaving 
space between them for the crank. Such engines usually employ 
splash lubrication for the crank pin and crosshead, that is, oil 
is placed inside the frame so that the crank can splash into it 
at each revolution and throw some of the mixture on the rubbing 
parts. This makes it necessary to shape the frame so as to 
contain the oil and also to cover the guides, connecting rod, and 
cranks to prevent the oil from being splashed out. The rectan- 
gular shape of the frame is well adapted to splash lubrication, 
as the frame itself makes a trough for containing the oil by merely 
extending the casting across the bottom. The cast-iron bottom 
of the frame serves not only to contain the oil but strengthens 
the frame laterally. Even when splash lubrication is not em- 
ployed, provision must be made for catching the oil that drips 
from the various bearings as it would soon destroy masonry 
foundations or soak into wooden floors around the engine. 

The cylinder of the high speed engine usually overhangs the 
frame, hence some provision must be made for fastening it to the 
frame. This is done by a ring of bolts at the crank end of the 

27 



28 STEAM ENGINES 

cylinder. These bolts are not used for aligning the cylinder with 
the frame and do not act as dowel pins; they are used only to 
hold the cylinder close up to the frame. The cylinder is aligned 
by means of a projection which fits accurately into a bored recess. 
In most cases the shoulder is on the frame and the recess is 
in the cylinder, but sometimes this arrangement is reversed, 
the shoulder being on the cylinder and the recess bored in the 
frame. The bolts for holding the cylinder may be outside the 
frame, or they may be inside and just at the end of the guides. 
The outside bolts are to be preferred because they are easier 
to reach in case the cylinder is to be removed for repairs. 

The frame of the Corliss engine has experienced decided 
changes in shape within recent years. Formerly the girder frame, 
as shown in Fig. 17, was most commonly used. The girder frame 
takes its name from the fact that the part of the frame between 
the main bearings and the cylinder does not rest directly on the 
foundation but acts as a girder, supported by a stand or legs 
placed about the middle of its length. The cross section of the 
girder frame, between the guides and the main bearings, is in the 
shape of the letter T to give it greater stiffness for the weight of 
metal in it. The guides form a part of the frame itself, being 
either bored to a circular shape or planed to a V-shape, hence this 
part of the frame is stiffened by the guides. The girder frame 
has the advantage of being light, and for small engines, particu- 
larly where the load is fairly uniform, it is very satisfactory; 
but for heavy and widely varying loads a more rigid and stronger 
frame is desirable. 

The latter class of service has brought into use on the larger 
sizes of Corliss engines the " heavy duty " frame, one form of which 
is shown in Fig. 18. This frame is built in one piece from the 
cylinder to the main bearings and is box shaped to allow it to rest 
squarely on the foundation throughout its length, but it is cut 
away on the outside from the guides towards the main bearing. 
The cylinder is bolted to the end of the frame and is supported 
independently. The guides are formed in the frame itself and 
the frame is formed into a complete circle at both ends of the 
guides to give greater strength and stiffness. In some heavy 
duty frames, the part forming the guides is made separate from 
the part containing the main bearing, being made in the shape 
of a barrel and bolted to the rear section and to the cylinder, but 
not resting directly on the foundation. 



PARTS OF THE STEAM ENGINE 



29 




30 



STEAM ENGINES 



The frames of vertical engines are usually A-shaped, as shown 
in Fig. 19. In the larger sizes they are made in two parts, the 
upper part or guide section being in one piece and bolted to the 
lower part or housing. The bottom of the frame is usually a 
rectangle in shape to permit it to rest squarely on the foundation. 
A web is cast entirely across the bottom to catch oil and to 
prevent it from soaking into the foundation. The cylinders are 
supported by the guide section and are bolted to it. In some 
marine engines, which are always of the vertical type, the A- 
shaped frame is modified, the cylinders being supported directly 
on steel columns and the guides bolted between them. This 
construction is used to secure lightness with strength. 

The Cylinder. — A 20'' X 24" cylinder of a locomotive is shown 
in Fig. 20. This one was chosen here because it is the cylinder 
of a plain slide valve engine, and, with its long ports and short 




Fig. 18. 

valve face represents one extreme; the short-ported Corliss 
engine is the other extreme. The cylinder body is a continuous 
shell except where the ports cut through it; and around each end 
there is a stout flange to which the cylinder heads are bolted. 

It will be noticed that the cylinder is bored out to two different 
diameters, the ends being from J^'' to }^" larger in diameter 
than the central portion. The central portion is called the 
''bore" and the larger end portion the ''counterbore." The 
length of the bore is such that the piston rings slightly overtravel 
it in order to prevent wearing shoulders in the bore. The 
counterbore is made large enough to allow reboring the cylinder 
two or three times to bring it back to a true cylindrical form after 
it has become worn by the piston. 

The heads are cast with a shoulder which is turned to a close 
fit with the counterbore, and they are then bolted on. The head 
end head is recessed to receive the nut which holds the piston on 
the piston rod, thus reducing the necessary clearance; and the 



PARTS OF THE STEAM ENGINE 



31 




Fig. 19. 



32 



STEAM ENGINES 



crank end is recessed on the outside to form a stuffing box. In 
this case, both heads are made thin, and ribs are cast on the 
outside of them to give greater strength. The walls of the 
cylinder are practically uniform in thickness, and, in order to 
accomplish this, recesses are left on the outside of the cyhnder 
where necessary. Uniform thickness of the walls of a steam 
cylinder is desirable in order to prevent unequal strains in the 
cyhnder, due to expansion caused by the heat to which it is 
subjected. 

That part of the casting forming the valve chest and steam 
passages is much more compHcated than the other parts of the 




Fig. 20. 

cyhnder on account of the necessary provision for fastening the 
cylinder to the saddle, and conducting the hve steam to it and 
the exhaust steam away from it. The valve seat consists of a 
raised rectangular table with the steam and exhaust ports leading 
up to its surface. The steam passage enters through the saddle 
and then divides, entering the bottom of the steam chest on each 
side of the valve seat at A and A. The exhaust passage Hes 
between the steam passages and it is of such width that a web B 
is cast across it to brace its side walls. 

The entire outside of the cylinder and valve chest is covered 
with planished sheet iron to give a smoother and neater appear- 



PARTS OF THE STEAM ENGINE 



33 



ance to it. The space around the barrel of the cyUnder, between 
the cyUnder and covering, is filled with nonconducting material 
such as asbestos, to reduce the loss of heat by radiation. 

The cylinder shown in Fig. 21 represents a usual type of con- 
struction for automatic high speed engines having flat valves. 
As compared with the cylinder shown in Fig. 20, this one has a 
larger diameter in proportion to its length, has shorter ports 
due to the greater width of valve, and the cylinder casting is 
somewhat simpler. The cylinder is counterbored, as is the case 
with all steam engine cylinders, and in this case the heads are 




Fig. 21. 

recessed to allow the ports to end behind the piston instead of 
ending flush with the cylinder walls. The heads are set into the 
cylinders and bolted on. On account of the different methods 
of fastening the piston to the piston rod, the head end head is 
not recessed but forms a plane surface to agree with the face of 
the piston. This head is made double with an air space between 
for insulation. The crank end head is recessed on the outside 
for the stuffing box, which extends sHghtly into the cyhnder. 
The piston is cored out to fit the projection on the head, and thus 
to reduce the clearance. The clearance is further reduced by 
the short ports. The exhaust port in this cylinder is wide and 
shallow, and its walls do not, therefore, require bracing. 



34 



STEAM ENGINES 



A Corliss engine cylinder is shown in Fig. 22. The bottom of 
the cylinder, which is rectangular in shape and flat, rests on its 
own bedplate and is bolted to it. The two admission valves are 
placed at the top of the cylinder and the two exhaust valves at 
the bottom, all being at the ends of the cylinder. The axes 
of the valves are placed across the cylinder, and the valve cham- 
bers are of rectangular form so that the cylinder has a square 
cornered appearance. In later types of engines the tendency 
is to round the valve chambers to conform to the shape of the 
valve, as shown in Fig. 11. The steam chamber S is cored out of 
the top of the cjdinder and extends all the way across it, giving 
a flat top to the cylinder. As this chamber contains steam at 
boiler pressure, webs are cast in it to give a greater strength. 




Fig. 22. 

The steam chamber is formed right upon the walls of the cylinder 
so it may act as a steam jacket to this part of the cylinder. 
The exhaust chamber, E, which contains steam at a lower tem- 
perature, is separated from the cylinder walls by a cored chamber. 
The exposed portions of the valve chambers are usually polished 
to decrease the radiation of heat, and the other parts of the 
cylinder are covered with nonconducting material. 

In the cylinder shown here the heads are cast very thin and are 
strengthened by webs cast on the outside of them, the head end 
head being covered by a cover plate to give a neater appearance. 
Both heads are cast with plain inner surfaces, and both faces 
of the piston are made flat to correspond. This allows the piston 
to be brought very close to the cylinder heads and thus reduce 
the clearance. The steam and exhaust ports are partly cored 



PARTS OF THE STEAM ENGINE 



35 



out of the heads so there may be some piston surface exposed to 
pressure when the piston is at the end of its stroke. 

The placing of the exhaust valves in a Corliss engine at the 
lowest point of the cylinder allows water to drain through them 
when the cylinder is being warmed up preparatory to starting 
the engine. In a slide valve engine the valve is placed at the 
side or on top of the cylinder, and drain valves or cocks must be 
placed in the bottom of the cylindero 

The cylinder of a four valve high speed engine, shown in Fig. 
23, is constructed the same as a Corliss cylinder, the principal 
difference being the greater diameter in proportion to the length. 




Fig. 23. 



The ports and valves of these engines are often made double in 
order to secure a greater opening with a small valve travel, 
which is desirable when the speed of the valve is high. The 
clearance in the cylinders of these engines is relatively greater 
than that in the Corliss engine because, while the piston may be 
brought as close to the head, the volume between the piston and 
head will be greater because of the larger diameter. In some 
makes of four valve high speed engines and also in some large 
Corliss engines an effort is made to reduce the clearance to a 
minimum by placing the valve chambers in the cylinder heads, 
as shown in Fig. 24; this reduces somewhat the length of ports and 
thereby reduces the clearance. 



36 



STEAM ENGINES 



The cylinder of a large marine engine is shown in Fig. 25. The 
peculiar feature about this cylinder is the shape of the heads and 
the fact that the cyhnder is fitted with a liner. The piston is 




Fig. 24. 



cone-shaped to conform to the shape of the heads, this shape being 
adopted to aid the drainage of water of condensation into the 
ports. Having a liner in the cylinder, as indicated at A , simpHfies 




Fig. 25. 



the work of casting the cylinder since the hner is cast separately, 
and permits of a sound and close grained hner being obtained while 
the cylinder proper is made of softer iron. The hner is fastened 



PARTS OF THE STEAM ENGINE 



37 



to the cylinder at the bottom by sunk head bolts in the inward 
projecting flange, the top being turned to a close fit. The ports 
may be either cut through the liner or carried around it; in this 
case they are carried around it. In order to secure lightness and 
strength, the heads are cast with double wall chambers which are 
strengthened by webs. This method of construction also reduces 
loss of heat from the cylinders since there is an enclosed air space 
next to the cylinder walls. 

The Piston. — An engine piston must meet the following 




Fig. 26. 

conditions: it must have enough strength to withstand the steam 
pressure acting upon it and yet be no heavier than necessary, 
as its weight controls its inertia and causes it to wear the cylinder, 
especially if the cylinder is horizontal; it must have a broad rim 
or working face, especially in horizontal engines, in order to have 
plenty of rubbing surface and reduce wear; it must be constructed 
to prevent the leakage of steam past it. The last result is secured 
by having the piston a little smaller than the bore of the cylinder 
and closing the gap between piston and cylinder by means of 
rings sprung into grooves in the piston. 

The box piston is by far the most common type used in cylin- 
ders up to 24 inches in diameter. One of the simplest forms of 



38 



STEAM ENGINES 



box piston is shown in Fig. 26. This piston consists of a simple 
hollow casting with flat faces of uniform thickness on both 
sides and with a hub cast into its center for receiving the piston 
rod. The piston is strengthened by casting webs across it so as 
to divide it into a number of compartments. A hole is cast into 
each of the compartments through which the core may be re- 
moved, after which it is drilled and tapped to receive a plug. 
In order to fasten the piston rod in the piston, the hub of the 
piston is bored smaller than the rod, and the end of the rod is 
turned to fit the hub which rests against the shoulder on the 
rod. The two are then fastened together with a countersunk 



A 




B 



Fig. 26a. 

nut screwed on the rod, the outer end of the nut being flush with 
the face of the piston. 

This piston is supphed with a single heavy packing ring placed 
at the center. It is common, however, to place two Kghter rings 
on the piston, one near each edge. In some cases the piston is 
supphed with four hght rings in two grooves placed near each 
edge. The rings are turned out of cast iron and are made a little 
larger in diameter than the bore of the cyhnder. They are 
then sawed through diagonally, only one cut being in a ring, and 
the ring is then sprung on the piston. The spring in the ring, 
since it is larger than the cyhnder, keeps it pressed outward against 
the walls of the cylinder and prevents steam from leaking past 
it. When a single ring is used, a chp is placed at the cut in the 
ring to prevent steam from leaking through, as shown in Fig. 
26a, but when the piston has more than one ring no chp is used 



PARTS OF THE STEAM ENGINE 



39 



the rings being simply placed on in such manner as to break joints. 
Instead of the faces of the box piston being fiat, they are often 





<..rD-.in,.^ 

B 

Fig. 27b. 



shaped to conform to projections on the inside of the cylinder 
heads. 

Two examples of locomotive pistons are shown in Fig. 27. In 
both of these designs a special effort is made to secure strength 
and Kghtness. The one shown at A is made 
entirely of cast iron of a simple T section and 
with two packing rings in the rim. This piston 
is forced on the tapered end of the piston rod by 
means of a nut. The piston shown at B is 
made in two parts, a central conical part made 
of steel, and a cast-iron rim which is bolted to 
the central portion by means of countersunk bolts. 
A peculiar feature of this piston is that the bottom, 
which carries the weight, has a much broader 
wearing surface than the top. This piston is 
pressed against a shoulder on the piston rod by 
means of a nut screwed on the end of the rod. 
The end of the rod has a slight taper to secure 
a better connection between the rod and the 
hub. 

The pistons of vertical engines, especially of 
the marine type, are made as light as possible 
consistent with proper strength. A wide rim or wearing surface 
is not necessary on these pistons, since the weight of the 
piston is not carried by the cylinder. One type of marine 
engine piston is shown in Fig. 28. This piston is constructed 




Fig. 28. 



40 



STEAM ENGINES 



with a steel web, and the entire rim is made of two cast-iron spring 
rings which are held in place by a follower ring bolted to the steel 
web. 

Corliss engine pistons are usually of the ''built up" type, con- 
sisting of several adjustable parts bolted together. A piston of 
this type is illustrated in Fig. 29. The ribbed body of this 
piston, in which the rod is fastened, is called a spider. The rim is 
made in two parts, 2 and 3, and is called a ''bull ring" or "junk 

ring." This carries a single heavy pack- 
ing ring. The bull ring is held in place 
by a follower plate, 4, which is fastened 
by tap bolts to the spider. The bull 
ring forms the wearing surface of the 
piston and can be adjusted by set screws 
so as to make the axis of the piston 
agree with that of the cylinder. The 

^ ■ ■.x.xx... Tx..^^ vx. bull ring is made in two parts so that 

M ' ' 1 ^ packmg rmg may be put m with- 

in ' ' /I ^^^ having to spring it over the piston 

and also because access may be had to 
the packing ring without removing the 
piston from the bore of the cylinder. 
The piston rod has a taper fit against 
a collar and is riveted over a heavy washer 
at the end, after which a key is passed 
through both rod and hub. 

Stuffing Box. — Stuffing boxes are used 
to prevent the leakage of steam at the 
points where the piston rod and valve rod 
pass through the cylinder head and valve chest, respectively. As 
the stuffing boxes at both of these points are constructed alike, a 
description of one will be sufficient. A stuffing box consists of 
two parts : first, an annular space surrounding the rod, as shown 
in Fig. 30; and, second, a cover plate called a "gland" extending 
into the stuffing box in such manner as to compress the packing 
material when it is screwed down. Two stud bolts are screwed 
into the cylinder head, one on each side of the stuffing box, and 
these extend through the gland and end in a nut so the gland 
may press upon the packing. The bottom of the stuffing 
box is most often cut away at an angle, as shown in Fig. 30, but 
sometimes it is made flat. The hole through which the rod 




Fig. 29. 



PARTS OF THE STEAM ENGINE 



41 



passes into the cylinder must be large enough to accommodate 
any lack of alignment between the rod and cylinder but must not 
be large enough to allow any of the packing material to be 
squeezed through it. 

Formerly, braids or strands of hemp soaked in tallow were 
wrapped around the piston rod and pressed into the stuffing box 
or packing, but, as increasing steam pressures and temperatures 
became common, this kind of packing became unsatisfactory. 
There are now a great variety of packings on the market made of 
vegetable fiber, asbestos, or rubber in various combinations, and 
frequently mixed with graphite for a lubricant, these being made 
of sizes and shapes to fit neatly into the stuffing box. Woven 





Fig. 30. 



packing is often made square in cross section and divided diago- 
nally into two parts so that when the gland is screwed down the 
packing is pushed out squarely against the rod. 

Metallic packing is often used in the stuffing boxes of steam 
engines, but requires a good alignment of the piston rod to work 
properly. Metallic packing is usually made in the form of 
babbitt metal rings, which are pressed against the piston rod, 
preventing the leakage of steam and causing but little friction. 
An example of this kind of packing is shown in Fig. 31. The 
gland, G, is merely a heavy cover plate made tight by a copper 
wire acting as a gasket. The ring 1 presses against the casing 
2 and forces the babbitt metal rings 3, 4, and 5 against the rod. 
These rings are made in segments and placed so as to break joints. 



42 



STEAM ENGINES 



A follower ring 6 is held in place by a heavy spring and keeps the 
rings in their proper position, but the spring is not depended upon 
to press the packing against the rod. This is done by the steam 
pressure acting behind the ring 6 so that the tightness of the 
packing varies with the steam pressure. 

The Crosshead. — The crosshead moves in a straight line 
between guides and is for the purpose of joining the piston rod to 
the connecting rod. It, therefore, has two joints : a stationary one 




Fig. 31. 



between the piston rod and crosshead; and a pin joint between 
the connecting rod and the crosshead so the connecting rod may 
be free to move. 

There are three general types of crossheads used upon station- 
ary engines, called respectively, the ^^wing," the ''block," 
and the ''slipper" crossheads. The wing crosshead is illustrated 
by Fig. 32. It consists of a heavy steel or cast-iron block forming 
three sides of a rectangle and having a heavy "wrist pin" passing 
between the side pieces or wings. The piston rod is threaded 
at the end and screwed into the front crossbar of the crosshead, 



PARTS OF THE STEAM ENGINE 



43 



being held securely by means of a lock nut. The wings form the 
rubbing surfaces, and, to reduce friction, they are often con- 
structed with grooves filled with babbitt metal. The guides consist 
of four flat bars between which the wings move. The guides are 
adjustable up and down to take up wear in the crosshead. The 
wrist pin is usually made of steel, separate from the crosshead, 
and is held in place by nuts on the ends, or sometimes by means of 




Fig. 32. 

a set screw. The wing crosshead is often used on the smaller 
sizes of engines, especially those of the plain sHde valve type. 

The block crosshead is most often found on Corliss and auto- 
matic high speed engines. As shown in Fig. 33, it consists of a 
heavy cast-iron block with the wrist pin passing through its 
center and with the piston rod screwed or keyed into the center 
at the front. The rubbing surfaces, located at top and bottom, 
are of circular shape for bored guides, and V-shaped for planed 
guides. Both top and bottom rubbing surfaces or ''shoes" 
are adjustable and have babbitt metal inserted in grooves or 



44 



STEAM ENGINES 



holes to reduce friction. With this type of crosshead the guides 
are not adjustable but the shoes of the crosshead are adjusted 
to take up wear and to bring the crosshead into alignment 
with the piston rod. The shoes are adjusted by means of 
wedges placed between the shoes and the body of the crosshead 
and moved by screws fitted with lock nuts to hold the wedge in 







Fig. 33. 



position after adjustment. The method of securing the wrist 
pin in this crosshead is shown by the taper ends of the pin com- 
bined with a nut on the end. The method of carrying oil to 
the rubbing surface of the pin is clearly shown. 




Fig. 33a. 

The slipper type of crosshead, Fig. 33a, resembles the wing 
type, but differs from it in having the wrist pin and main body 
of the crosshead placed above the wings. In this case the wings 
are comparatively thin but the rubbing surface, which is all of the 
bottom of the crosshead, is broad. The guides consist of a 
flat planed surface on the engine frame, with a rectangular bar 



PARTS OF THE STEAM ENGINE 



45 



at each side to fit on top of the shpper. These rectangular 
bars are adjustable to take up the wear of the crosshead. The 
wrist pin consists of a simple steel cylinder and is clamped 
between the jaws of the crosshead, which are split and provided 
with bolts for this purpose. The wrist pin is arranged so it may 
be turned through 90 degrees as it wears, thus keeping it round. 

Connecting Rods. — The connecting rod connects the wrist 
pin and crank pin and serves to transmit the force acting upon the 
piston to the crank pin. For one-half of a revolution of the fly- 
wheel the forces acting along the connecting rod are. pushing 
and for the other half of the revolution they are pulling. The 
connecting rod consists of an adjustable bearing at each end 
connected by the shank or rod proper. 

The cross section of the rods is made in various shapes, depend- 
ing upon the type of engine with which they are to be used. For 
slow running engines of the Corliss type the rod is usually round. 





Fig. 34. 



being largest at the middle and tapering towards the ends. 
With engines of higher speed the rod is often shaped like a long 
cone, tapering towards the crank end and flattened on the sides 
so as to approach a rectangular cross section as the diameter 
increases. High speed stationary engines and some locomotives 
have a rod of rectangular cross section increasing in depth 
towards the crank end. Many of these engines have the cross 
section of the rod of an I-shape in order to make them light and 
strong. Marine engines usually have round rods. 

Great variety is found in the construction of the ends of the 
connecting rod, but they will usually fall in one of the three 
general classes called respectively the "box" or ''solid end," 
the ''strap end," or the "marine end." Often the rod will have 
one end of the "box" type and the other end of the "strap" 
type. The box or solid end type of construction is well illustrated 
in Fig. 34, which shows that the end of the rod is flattened and 
has a rectangular slot milled into it. Into this slot are placed 
the two halves of the bearings, which are made of brass or bronze. 



46 



STEAM ENGINES 



These halves are separated by a small space to allow them to 
be brought closer together as they wear away. The ''brasses" 
are cast with flanges at the sides which fit the sides of the rod to 
prevent their movement sidewise. Behind one of the brasses 
is placed a wedge-shaped block with a screw held by nuts at 
each end passing through it. Adjustment of the brasses is made 
by turning this screw and moving the wedge-shaped block up- 
ward, thus forcing the brasses closer together. By having one 
of the adjusting wedges placed on the inside of the end and the 
other placed on the outside of the end, the length of the connect- 
ing rod, which is the distance from the center of the wrist pin 
to the center of the crank pin, is not changed when both ends 
are adjusted at the same time. 

A strap end connecting rod is illustrated in Fig. 35. In this 
form, the connecting rod ends at the brasses and a separate steel 




Fig. 35. 

strap passes around the brasses and laps over the end of the rod 
at top and bottom. The strap is fastened to the connecting 
rod by means of two bolts which pass entirely through both 
ends of the strap and through the connecting rod and are secured 
by a nut and lock nut. Keys are inserted between the strap and 
connecting rod at top and bottom to keep the strap in line with 
the rod and to relieve the bolts of shear and permit the use of 
lighter bolts. The brasses are adjusted by means of a wedge in 
back of one of the brasses, the wedge being moved up or down by 
a bolt threaded into the brass and passing through the strap with 
a lock nut on the outside. 

Strap end connecting rods are used commonly on locomotives 
but are dropping out of use on stationary engines. When used 
on locomotives, the brasses form the bearing against the pins, 
but for stationary engines the brasses are usually lined with 
babbitt metal. Locomotive connecting rods have the brasses 
adjusted by means of a wedge driven down behind one side of 
the brass and locked in place by means of a set screw. 

The right hand end of the connecting rod shown in Fig. 36 



PARTS OF THE STEAM ENGINE 



47 



is of the marine type; the left-hand end is of the box type, 
described before. Marine end connecting rods are used on all 
marine engines and on some vertical stationary ones. The 
example shown here is from a stationary engine. Those designed 
for marine engines usually have both ends of the connecting 
rod of the marine type, instead of only one. In the marine end 
connecting rod, adjustment of the brasses is secured by means 




Fig. 36. 

of bolts placed parallel to the connecting rod and passing through 
shoulders on the rod and on the removable end, as shown. This 
form of adjusting device permits a short end for the rod, which 
accounts for its common use on marine engines, where the crank 
pin passes close to the floor. 

Crank and Crank Pin.^ — The steam pressure acting upon the 
piston is transmitted through the connecting rod to the crank 
pin and then through the crank to the shaft. Engines may be 
divided into overhung crank engines, in which the crank is at 
the end of the shaft; and into center crank engines, in which 
the crank is placed at or near the center of the shaft. 

An overhung crank for 
an engine of medium size 
is shown in Fig. 37. This 
crank is made in the form 
of a cast-iron disk, with 
holes bored to receive the 
crank pin and the shaft. 
The crank disk is made 
thicker opposite the crank 
pin than it is on the 
crank pin side in order to counterbalance it. The crank disk 
is either forced on the shaft and keyed, as shown here, or 
else shrunk on, in which case a key is unnecessary. The crank 
pin is usually forced in by hydraulic pressure and then riveted 
over. Overhung cranks for large slow speed engines are some- 
times forged from steel in the shape shown in Fig. 38. This 




Fig. 37. 



48 



STEAM ENGINES 



shape does not permit of as much counterbalancing as the disk 
shape, but, on the other hand, slow speed engines do not require 
as heavy counterweights as do high speed engines. 

A crank for a center crank engine is shown in Fig. 39. This 





i I I i _F' 



Fig. 38. 

consists of two cast-iron disks with counterweights, fastened to 
the shaft, which is made in two sections, and with the crank 
pin connecting the two disks. Another form of center crank is 
illustrated in Fig. 40, in which the cranks, crank pin, and shaft 
are all in one piece and forged from steel. The counterweights 

are made of cast iron and 
bolted to the cranks oppo- 
site the crank pin, as shown. 
Bearings. — There are al- 
ways two bearings for the 
shaft on an engine. In 
center crank engines these 
two bearings are ahke, but 
in side crank engines the 
one next to the crank, called the main bearing, is much heavier 
than the other, or outer bearing. 

The bearings must not only support the weight of the shaft 
and flywheel, but must resist the thrust of the piston, and also 
resist the pull of the belt, if there is one. The resultant of all 
of these forces causes the bearings to wear at an angle with the 




Fig. 39. 



PARTS OF THE STEAM ENGINE 



49 



horizontal rather than at the bottom and top. Since the wear 
comes at an angle to the horizontal, provision must be made for 
adjusting the bearings at an angle. This is done by dividing the 
bearing into two parts with the division line between the parts 
making an angle with the horizontal, or by dividing the bearing 
into four parts so that adjust- 
ment may be made at the 
side or at the top. 

Two part bearings, such as 
are often found on high 
speed engines, are illustrated 
in Figs. 41 and 42. In Fig. 41 
the bearing is divided along a 
line making about 45 degrees 
with the horizontal. The lower part has a flat plate cast on its 
outer surface which rests in the frame and which may be adjusted 
vertically by shimming, or placing thin sheets of metal under it. 
The top half has one flat plate at the side and another at the top. 
This part of the bearing may be adjusted horizontally by means 
of bolts passing through the back of the frame and bearing against 




Fig. 40. 




Fig. 41. 

the flat plate at the side. The cap rests on the flat plate at the 
top and is held in place by bolts passing down into the frame. 
The bearing is lined with babbitt metal having a series of 
diagonal grooves cut in it for distributing the oil throughout the 
length of the bearing. 

The bearing shown in Fig. 42 is similar to the above except 
for the method of adjusting. In this bearing the top part is 



50 



STEAM ENGINES 



moved horizontally by means of a wedge placed behind it and 
which may be moved up or down by a bolt extending through the 
cap. 




Fig. 42. 



Large horizontal engines usually have main bearings of the type 
shown in Fig. 43. In this type the bearing is divided into four 
parts, the side pieces being adjustable. One of these side pieces 




Fig. 43. 



is adjusted by shims and the other by set screws passing through 
the frame. The top part of the bearing is made as a part of the 



PARTS OF THE STEAM ENGINE 



51 



cap, which is held in place by bolts passing down into the frame 
and which serves to hold the parts of the bearing together. The 
bearing is lined with babbitt metal with diagonal grooves cut in 
it. The advantage of this type of construction is that the bearing 
may be removed without removing the shaft, by taking off the 
cap, slightly lifting the shaft, and turning the bearing around. 

The Flywheel. — The flywheel serves a threefold purpose: 
It sometimes serves to transmit the power of the engine to other 
machines by means of belts; to store up enough energy near the 
middle of the piston stroke to carry the engine past center; and, 
by storing up energy at one part of the stroke and giving it out 



^^^^^^^^#^^^^^^ 





> 



iU 

4^-i-m 




Fig. 44. 



again at other parts, to prevent fluctuations of speed during a 
revolution of the flywheel. 

Small engines are usually supplied with two ordinary belt 
wheels. In slow speed engines one of these wheels is sometimes 
larger than the other, but in medium and high speed engines 
both wheels are of the same size and kind. Even when high speed 
engines are direct connected to generators there is often one 
flywheel. 

Wheels less than 9 feet in diameter are usually cast in one 
piece, but with the hub split on one side, as shown in Fig. 44, so 
it may be clamped to the shaft by two bolts, one on each side of 
the spokes. These bolts are not depended upon to hold the 
wheel, however, but merely to simplify putting the wheel on the 



52 



STEAM ENGINES 



shaft. The hub is held securely to the shaft by means of a close 
fitting key. 

Flywheels between 9 and 16 feet in diameter are commonly 
made in halves, and larger sizes are divided into a greater number 
of parts, as a 16 foot piece is about as large as can be shipped on an 
ordinary flat car. Fig. 45 illustrates the method of joining the 
halves of a large flywheel. The hub is clamped to the shaft by 




Fig. 45. 

bolts extending all the way through on each side of the shaft, after 
which the hub is keyed to the shaft. This type of wheel is used 
only as a flyw^heel and not as a belt w^heel, hence, the rim is made 
narrow and deep in order to concentrate its weight as far from the 
center of the shaft as possible. The rim is fastened together by 
inwardly projecting flanges bolted together and by I-shaped bars 
or Hnks let into the sides of the rim. These hnks are machined 
exactly to length between heads and the slots similarly machined. 
The Hnks are made shorter than the slots by from one in one 
thousand to one in eight hundred. The Hnks are then heated 
and expanded until they will go into the slots. Upon cooling, 
the hnks contract and hold the halves of the rim together tightly. 



CHAPTER IV 
HEAT, WORK, AND PRESSURE 

Force. — If a weight is held in the outstretched hand, a down- 
ward pull on the hand will be experienced. Unless this pull is re- 
sisted, the hand will be moved downward. The pull on the hand 
in this case is called a force. A force always tends to produce 
motion and it may, therefore, be defined as that which produces 
motion or tends to produce motion. If a force is applied to any 
stationary object, the object will move unless the force is resisted 
or opposed by another force equal in amount but opposite in 
direction. If a moving object is not acted upon by any force, 
or is acted upon by forces which are equal in amount but opposite 
in direction, the object will continue to move with a uniform 
velocity; but if the forces acting upon the moving object are not 
equal in amount and opposite in direction, the velocity will 
increase as long as such forces are applied. 

In a steam engine, the piston is acted upon by the force of the 
steam pressure which causes the piston to move, and the motion 
is transmitted to the flywheel. The velocity of the flywheel will 
increase until the forces opposing its motion are equal in amount 
to those causing it to move, after which it will move with uniform 
velocity. 

A force is measured by the number of pounds with which it 
pulls; thus a force of 10 pounds is the downward force exerted 
by a weight of 10 pounds, or the force necessary to lift a weight of 
10 pounds. 

A force acting upon the crank pin or rim of the flywheel pro- 
duces what is called a torque or twisting moment. A torque can- 
not be measured in pounds because its amount depends both upon 
the force applied at the circumference of the circular path of the 
force and also upon the distance from the center to the point 
at which the force is applied. In the case of the engine crank, 
the torque depends both upon the actual force applied to the 
crank pin and upon the length of the crank from the center 
of the shaft to the center of the crank pin. The amount of the 
5 53 



54 STEAM ENGINES 

torque is expressed in foot-pounds and is equal to the number of 
pounds of force applied at the circumference multiplied by the 
radius of the circular path, in feet. If a force of 1000 pounds is 
applied to a crank pin which is two feet from the center of a shaft, 
the torque or twisting moment will be 

1000 X 2 = 2000 foot-pounds 

Work. — When force is applied to an object in such a manner 
as to cause the object to move, work is performed. If the object 
does not move, no work is performed no matter how large the 
applied force may be. Steam admitted behind the piston of a 
steam engine causes the piston to move, hence the steam performs 
work upon the piston. If the piston is blocked so that it cannot 
move and steam is admitted behind it, no work will be performed 
because, while the steam pressure may be as great as before, 
the piston in this case does not move. 

The amount of work performed is always equal to the force 
applied, expressed in pounds, multiplied by the distance through 
which the object moves, expressed in feet. The unit of work is 
the foot-pound and is the amount of work done when a force 
of one pound moves through a distance of one foot. 

Example. — How much work is performed when a steam pressure of 80 
lb. per sq. in. acts upon the piston of a steam engine which is 18 inches 
in diameter and which moves it through a distance of two feet? 

Solution. — The area of the piston is 

.7854 X 182 = 254.5 sq. in. 

The force acting on the piston is 

254.5 X 80 = 20,3601b. 

The work performed when this force moves through a distance of two 
feet is 

20,360 X 2 = 40,720 ft.-lb. 

Work and torque should not be confused with each other, even though 
both are expressed in foot-pounds. In the case of work there is motion; 
in the case of torque there is no motion. 

Energy.- — Energy is the ability to do work. Water falling 
over a waterfall is able to perform work, hence falhng water 
possesses energy. Steam performs work in a steam engine, 
hence steam possesses energy. Anything which is capable of 



HEAT, WORK, AND PRESSURE 55 

performing work, or producing motion by overcoming a force, 
possesses energy. 

There are several kinds of energy such as mechanical energy or 
the energy of motion, electrical energy, heat or thermal energy, 
and chemical energy; and any of these different kinds of energy 
may be changed into any other kind. In a power plant the coal 
which is burned under the boilers contains chemical energy, 
due to the various chemical substances of which it is composed. 
When the coal is burned, its chemical energy is changed into heat 
energy. The heat energy causes the water in the boiler to form 
steam and the steam containing the heat energy is carried to the 
cylinder of a steam engine. The heat energy which the steam 
contains is changed, in the cylinder of the steam engine, into the 
mechanical energy of the moving piston. The mechanical energy 
of the moving piston is transferred through the connecting rod 
and crank to the shaft. A direct-connected generator on the 
shaft changes the mechanical energy into electrical energy and 
this electrical energy may be changed again into mechanical 
energy by means of a motor. 

Energy cannot be measured directly but it may be measured 
by the effects which it produces. The different kinds of energy 
mentioned above produce different effects, hence each kind has a 
different unit which is based upon the effects produced by this 
kind of energy. The unit for mechanical energy, or the energy 
of motion, is the foot-pound the same as for the unit of work, 
since the effect of mechanical energy is to produce work. Any 
object which is capable of performing one foot-pound of work is 
said to contain one foot-pound of energy. 

Heat.' — In the study of steam engines we are concerned prin- 
cipally with heat energy, or simply heat, as it is commonly called, 
since the work performed in the cylinder of a steam engine 
comes from the heat which is contained in the steam. 

Both heat energy and mechanical energy are energies of motion 
but there is this difference between them; that mechanical energy 
is a motion of a body taken as a whole while heat energy is a 
motion of the particles of which a body is composed, aside from 
any motion which the body as a whole may have. 

All substances are composed of very small particles called 
molecules, which are so small that they cannot be seen, even by 
the aid of the most powerful microscope. The molecules are in 
constant motion, vibrating back and forth, yet held together 



56 STEAM ENGINES 

by a force of attraction which they have for each other and which 
keeps them vibrating within certain Hmits. Since the molecules 
are too small to be seen, and vibrate through a very short dis- 
tance, they apparently cause no motion of the body as a whole, 
even though they are vibrating back and forth among themselves 
at a very rapid rate. The energy of these vibrating molecules 
is heat energy, or what is called heat. 

If a nail is placed on an anvil and struck rapidly with a hammer, 
the nail becomes hot or its temperature is increased, and, if the 
blows are struck faster, the temperature of the nail becomes 
higher; that is, the temperature of the nail depends upon the 
rapidity of the hammer blows. In vibrating back and forth, 
the molecules of a substance strike each other a great many 
blows, hence the temperature of the substance depends upon 
the rapidity with which the molecules vibrate, in the same way 
that the temperature of the nail in the above example depends 
upon the rapidity of the hammer blows, and the faster the mole- 
cules vibrate the higher the temperature of the substance will be. 
When heat is applied to a substance, the molecules receive 
energy and vibrate faster, hence, heating a substance increases 
its temperature. 

A substance may exist as a solid, a liquid, or a gas, depending 
upon its temperature and the amount of heat which it contains. 
At the lower temperatures a substance will be in the form of a 
solid; as it is heated to higher temperatures, it changes into a 
liquid; and as it is heated to still higher temperatures, it changes 
to a gaseous state. When a solid substance is heated, its mole- 
cules vibrate faster and faster and the temperature of the sub- 
stance increases. The faster the molecules vibrate the longer 
their path tends to be, hence, the more nearly they come to break- 
ing down the attraction which the molecules have for each other 
and which preserves the shape of the solid substance. But this 
partial breaking down of the force of attraction between mole- 
cules allows the substance to increase in size. This explains 
the expansion of substances when they are heated. 

As the solid substance is heated to a higher temperature, the 
vibration of the molecules becomes so fast and the blows which 
they strike become so hard that the attraction between the mole- 
cules is partially broken down and the substance takes a liquid 
form, in which it does not retain a definite shape, but the mole- 
cules are free to move with respect to each other. Even in the 



HEAT, WORK, AND PRESSURE 57 

liquid form there is a certain amount of attraction between the 
molecules; enough to keep them in one body. 

As heat continues to be applied to the hquid, its molecules 
vibrate faster, it increases in temperature, and continues to 
expand. Some of the molecules near the surface of the liquid 
vibrate with enough force to break through the surface and then 
go into the space above the liquid; some of these returning, 
others remaining in the outer space. This is the effect known as 
evaporation. Finally the temperature of the liquid becomes so 
high that great numbers of the molecules pass into the space 
above the liquid and boiling begins. 

The only difference between evaporation and boiling is that 
evaporation takes place only at the surface while in boiling the 
liquid is changed into a vapor both at the surface and in the body 
of the liquid, usually along the surface through which the heat 
passes into the liquid. This part of the vapor is formed in 
bubbles, which, being lighter than the liquid surrounding them, 
rise to the surface and burst. In a vapor or gas the attraction 
between the molecules has been completely broken down and they 
are free to move anywhere within the vessel which contains 
them, hence, a gas always expands and fills completely any 
vessel in which it is placed. The molecules of a gas are vibrating 
rapidly and, as there is no attraction between them, they move 
in straight lines from one end or side of the vessel to the other. 
In this way they are continually striking blows against the sides 
of the containing vessel. Just as the particles of water in a 
stream from a hose exert a pressure upon any object which 
the stream strikes, so the blows of the molecules against all 
sides of a vessel produce a pressure upon them and cause the gas 
to expand if the vessel is enlarged. 

Temperature. — Temperature should not be confused with 
heat. Temperature is only one of the effects of heat, and is not 
heat itself. Temperature is a measure of the rapidity of vibra- 
tion of the molecules and, therefore, is a measure of the intensity 
of heat. 

Unit of Heat. — Heat energy cannot be measured directly but is 
measured by its effects. The most common effect of heat is 
increasing the temperature of a substance, and as water is one of 
the most common substances, the effect of heat in raising the 
temperature of water is used in measuring quantity of heat. The 
unit of heat is called the British Thermal Unit (abbreviated 



58 STEAM ENGINES 

B.t.u.) and is taken as the quantity of heat which is required 
to raise the temperature of one pound of water from 62° to 63° F., 
this temperature being chosen because the amount of heat 
required to change the temperature of one pound of water one 
degree varies sUghtly at different temperatures. However, 
this variation is so shght that it may be neglected for most 
practical purposes, and the unit of heat taken as the quantity of 
heat which is required to raise the temperature of one pound of 
water one degree, without reference to any particular temperature. 

Mechanical Equivalent of Heat. — Since heat energy is capable 
of performing work there must be a numerical relation between 
heat and work. It has been found by experiment that one 
British Thermal Unit (B.t.u.) is equivalent to 778 foot-pounds of 
work and, therefore, also to 778 foot-pounds of mechanical energy. 

Specific Heat. — Experiment shows that different substances 
require different amounts of heat to raise their temperature 
one degree. Thus, one pound of cast iron requires .1189 B.t.u. 
to raise its temperature one degree, while one pound of lead 
requires .0305 B.t.u. to raise its temperature one degree. The 
number of heat units required to raise the temperature of o-ne 
pound of any substance one degree is called the specific heat of 
that substance. On this basis, the specific heat of water is one. 

Power. — Power is the rate of doing work. Work and power 
should not be confused. Work does not take into account the 
length of time required to perform it, while power does. In order 
to raise a weight of 4400 pounds through a distance of 60 feet, 
the amount of work required is 4400 X 60 = 264,000 ft.-lbs., 
and this amount of work is the same whether the weight is 
lifted in one minute or in one hour, but the power required to 
raise the weight will be greater for the shorter time in which the 
weight is raised than for the longer time. 

The unit of power adopted in engineering work is called the 
horsepower (abbreviated Hp.) and is the performance of 33,000 
foot-pounds of work in one minute. This is equivalent to the 
performance of 550 foot-pounds of work in one second, or of 
1,980,000 foot-pounds in one hour. The 264,000 foot-pounds 
of work mentioned in the example above would require : 

OQ Lf.^ = 8 horsepower if performed in one minute, or 

00}\J\J\J 

' = 0.133 horsepower if performed in one hour. 



1,980,000 



HEAT, WORK, AND PRESSURE 59 

Since 778 foot-pounds of work are equivalent to one British 

33 000 
Thermal Unit, one horsepower is equivalent to = ' = 42.42 

B.t.u. per minute, or 
1,980,000 



778 



= 2545 B.t.u. per hour, in round numbers. 



Example. — In a certain power plant 400 pounds of coal are burned each 
hour. The coal has a heating value of 12,000 B.t.u. per pound, of which 
the engines utilize 10 per cent. How much power is developed by the 
engines? 

Solution. — Heat liberated by the burning coal per hour = 400 X 12,000 = 
4,800,000 B.t.u. Heat utihzed by the engines per hour = 4,800,000 X .10 
= 480,000 B.t.u. Horsepower equivalent of 480,000 B.t.u. per hour = 
480,000 



2545 



= 188.6 Hp. 



Atmospheric Pressure. — The earth is surrounded by a body of 
air which exerts a pressure upon everything upon the surface of 
the earth, and this pressure must be taken into account in nearly 
all calculations dealing with pressure. The pressure exerted by 
the air is due to its w^eight and amounts to 14.7 lb. per sq. in. 
at sea level. If the atmospheric pressure is measured at a point 
above sea level, as on a mountain, it will be less than 14.7 lb. per 
sq. in., because the weight of air above this point is less. The 
following table shows the atmospheric pressure at various eleva- 
tions above sea level: 



Elevation above 


Atmospheric pressure 


sea level in feet 


lbs. per sq. in. 





14.70 


1,000 


14.20 


2,000 


13.67 


3,000 


13.16 


4,000 


12.67 


5,000 


12.20 


6,000 


11.73 


7,000 


11.30 


8,000 


10.87 


9,000 


10.46 


10,000 


10.07 



Besides varying with the elevation above sea level, the atmos- 
pheric pressure also varies slightly with the weather, but the 
variation from this cause is not very great. 

Vacuum. — In engineering work a vacuum is a space in which 
the pressure is less than atmospheric pressure. An absolute 



60 STEAM ENGINES 

vacuum is a space in which there is no pressure. It is almost 
impossible to produce an absolute vacuum, hence in most 
engineering work we have to deal with a partial vacuum, in 
which there is some pressure although not so much as 
atmospheric pressure. 

Barometer. — The atmospheric pressure may be measured by 
an instrument called a barometer. A simple form of barometer 
may be constructed as follows : A glass tube about 32 inches long, 
closed at one end, is completely filled with mercury. The finger 

tB is then held over the open end of the tube to prevent 
the mercury from spilling and the tube is quickly 
I inverted and the open end placed in a cup of mercury, 
as shown in Fig. 46. If the finger is then removed 
from the end of the tube, the mercury will sink in the 
tube until it stands at a certain height H above the 
surface of the mercury in the cup, depending upon 
the atmospheric pressure. 

The space above the mercury in the tube is as near 

! a perfect vacuum as can be produced. The full atmos- 

I pheric pressure acts upon the surface of the mercury 

i in the cup, hence the mercury in the tube will stand 

i at such height that its weight will just balance the 

I pressure of the atmosphere on an area equal to that of 

i the cross section of the tube. Since a cubic inch 

L |7 — ^ of mercury weighs .4908 pound, the height of the 

^^ mercury in the tube, in inches, multiplied by .4908 

^^' ' gives the atmospheric pressure in pounds per square 

inch. Thus, if the height H of the mercury is 29 inches, the 

atmospheric pressure is 29 X .4908 = 14.23 lbs. per sq. in. 

Absolute and Gage Pressures. — ^Gages which are used for 
indicating steam pressure are constructed so they read the 
amount of pressure above that of the atmosphere; that is, 
they do not read the true pressure or pressure above an absolute 
vacuum, but instead have their zero point at the atmospheric 
pressure. For this reason it is necessary to add atmospheric 
pressure to that indicated by the gage in order to find the true 
pressure or pressure above an absolute vacuum. The pressure 
indicated by a gage is called gage pressure and pressure measured 
above an absolute vacuum is called absolute pressure. The abso- 
lute pressure is equal to the atmospheric pressure plus the gage 
pressure. 



HEAT, WORK, AND PRESSURE 



61 



Example. — At a certain place in which the height of the mercury in a 
barometer stands at 28.5 inches, the steam gage on a boiler reads 110 lbs. 
per sq. in. What is the absolute pressure of the steam in the boiler? 

Solution. — The atmospheric pressure equals 

28.5 X .4908 = 13.98 (practically 14.0) lbs. per sq. in. 

The absolute pressure in the boiler equals 

14 + 110 = 124 lbs. per sq. in. 

When the number of pounds pressure is given without stating whether 
it is gage or absolute pressure, it is usually understood to be gage pressure. 

Measuring Vacuum. — A vacuum is measured by attaching 
it to a mercury column, similar to a barometer, and reading the 

n 



g^^zzg 





Fig. 47. 

number of inches of mercury which it will support. If the top 
end of the glass tube in Fig. 46 is opened so the atmospheric 
pressure can act on the surface of the mercury in the tube, the 
mercury will immediately fall to the same level as that in the 
cup. If, now, the top of the glass tube is attached to an air 
pump, as shown in Fig. 47, and the pressure in the tube reduced 
below that of the atmosphere, then the mercury will rise in the 



62 



STEAM ENGINES 



& 




E 5 




Fig. 48. 



tube to a height which 
balances the difference 
to%oB^li^^ ''" ^^ pressure between the 

atmosphere and that in 
the pump. For example, suppose the height 
of mercury in the tube attached to the 
pump in Fig. 47 is 6 inches. The reduction 
in pressure in the tube then amounts to 
16 X .4908 = 7.85 lbs. per sq. in 

The actual pressure above the mercury 
would be the difference between the atmo- 
spheric pressure and 7.85 lb. per sq. in. If 
the atmospheric pressure were 14.7 lb. per 
sq. in., the actual pressure above the mer- 
cury would be 

14.7 - 7.85 = 6.85 lb. per sq. in. 

It is usual to read and express the amount 
of vacuum in terms of inches of mercury 
thus: the amount of vacuum in the above 
example would be called ''16 inches of 
vacuum." 

A mercury gage for measuring vacuum 
may be constructed as follows: A glass 
tube about 68 inches long is bent into a 
U-shape as shown in Fig. 48, and is attached 
to a board with a scale divided into inches 
and tenths of an inch placed between the 
legs of the U-tube. Mercury is poured into 
the glass tube until the legs are about half 
full, and the instrument is then ready to 
use. As long as atmospheric pressure acts 
upon the surface of the mercury in both 
legs of the U-tube, the mercury in these 
legs will stand at the same height, but if 
one leg is attached to a vessel in which 
there is a vacuum, as, for example, the 
condenser of a steam engine, then the mer- 
cury will rise in this branch and fall an 
equal amount in the other branch. The 



HEAT, WORK, AND PRESSURE 63 

difference in height of the two columns of mercury will then 
represent the difference in pressure between the atmosphere and 
that in the condenser. For example, suppose the difference in 
height of the two columns of mercury is 16 inches. The re- 
duction in pressure in the leg attached to the condenser is 

16 X .4908 = 7.85 lbs. per sq. in. below that of the 

atmosphere. If the pressure of the atmosphere is 14.7 lbs. per 
sq. in., the actual pressure in the condenser will be 

14.7 - 7.85 = 6.85 lbs. per sq. in. 

The amount of vacuum in this example would be called ''16 
inches of vacuum." 

Another form of vacuum gage, and the one most commonly 
used, is constructed like a pressure gage, except it reads pressures 
below atmosphere instead of above, and instead of reading in 
pounds per square inch it reads in inches of mercury; its readings, 
therefore, have the same meaning as those of the two vacuum 
measuring devices previously described. 

It should be noted that the same amount of vacuum as indi- 
cated by a vacuum gage does not always mean the same thing. 
Thus, in New York, which is at sea level, a vacuum gage on a 
condenser which reads 22 inches of vacuum shows that the abso- 
lute pressure in the condenser is 

14.7 - (22 X .4908) = 4.1 lbs. per sq. in., 

while in Butte, Montana, which is approximately 5000 feet above 
sea level and where the atmospheric pressure is about 12.2 lbs. 
per sq. in., a vacuum of 22 inches would show that the absolute 
pressure in the condenser is only 

12.2 - (22 X .4908) = 1.4 lb. per sq. in. 



CHAPTER V 
PROPERTIES OF STEAM 

Formation of Steam. — Suppose we have an open pan contain- 
ing water at a temperature of 32° and that the atmospheric 
pressure is 14.7 lb. per sq. in. If the pan is placed over a fire, 
the temperature of the water begins to rise as soon as the water 
absorbs heat. The temperature of the water will continue to 
increase until it reaches 212° when small bubbles of steam will 
begin to form at the bottom of the pan and rise to the surface 
of the water and burst, liberating the steam which they contain. 
It will be noted that the temperature of the water does not rise 
above 212° even though we continue to apply heat to it. If we 
apply heat faster, the boiling occurs more rapidly. If we apply 
heat slower, the boiling occurs slower, but the temperature remains 
constant at 212°. If we place thermometers in both the water 
and the steam rising from the water, both of them will indicate 
212°, showing that the temperature of the steam is the same as 
that of the water with which it is in contact. The temperature 
of the steam cannot be raised above that of the boiling water 
as long as it remains in contact with the boiling water, since any 
attempt to do so will simply result in boiling the water faster. 
The steam may, however, be collected and removed from the 
presence of water and then heated to a higher temperature. 

If we should measure the amount of heat supplied to the water 
while its temperature was increasing from 32° to 212°, we would 
find that 180 B.t.u. had been supplied for each pound of water 
in the pan. If we should measure the amount of heat supplied 
to the water after it had reached 212°, we would find that 970.4 
B.t.u. had been supplied for each pound of steam formed. We 
would also find, upon measurement, that while one pound of the 
water occupies a volume of only .0155 cubic feet, the volume of 
the steam formed from it occupies a volume of 26.79 cubic feet 
or about 1700 times the volume of the water from which it was 
formed. The large quantity of heat absorbed by the water after 
it has reached the boiling temperature is utilized in breaking 

64 



PROPERTIES OF STEAM 65 

down the attraction of the molecules for one another and in 
increasing the volume of the steam from that of the water to that 
of the steam, this increase of volume taking place against the 
pressure of the water and the pressure on the surface of the water, 
which, in this case, is the atmospheric pressure of 14.7 lb. per 
sq. in. 

If the water, in the above example, had been heated in a closed 
vessel under a pressure greater than 14.7 lb. per sq. in., the tem- 
perature at which boiling occurred would have been greater 
than 212°, and if the pressure in the vessel had been less than 
14.7 lb. per sq. in., the temperature at which boiling occurred 
would have been less than 212°, showing that the temperature 
at which water boils depends upon the pressure acting upon the 
surface of the water. 

In the case of the water heated in an open pan, mentioned 
above, it is to be noted that the total amount of heat required 
to form one pound of steam may be divided into two distinct 
parts; first the amount of heat absorbed by the water in raising 
its temperature from 32° to the boiling temperature, and, second, 
the amount of heat required to change the water into steam after 
it has reached the boiling temperature. The first of these is 
called the ^'heat of the liquid^^ or sometimes the "sensible heat 
of the steam,'' because this part of the heat is sensible to the touch 
or affects a thermometer. The second part of the heat mentioned 
above is called the "latent heat'' or the "latent heat of evaporation," 
because it is not sensible heat but is latent. In the case of water 
heated under a pressure of 14.7 lb. per sq. in. the heat of the 
liquid amounts to 180 B.t.u. and the latent heat to 970.4 
B.t.u. The sum of these two quantities or 180 + 970.4 = 
1150.4 B.t.u. is the total quantity of heat supplied in forming 
one pound of steam from one pound of water at an initial temper- 
ature of 32°. The name of "total heat" is given to this quantity. 

Steam Tables. — The boiling temperature, heat of the liquid, 
latent heat, total heat, and volume of steam formed at various 
pressures have been found from experiment, or calculated, and 
this information is placed in tables, called steam tables, one of 
these tables being found at the end of this chapter. It will be 
noted that this table is headed ''Properties of Saturated Steam." 
By saturated steam is meant steam which has the same temperature 
at which it was formed. Steam in contact with the water from 
which it was formed is saturated steam. The various quantities 



66 STEAM ENGINES 

given in the steam table are for one pound weight of dry steam, 
hence to find these quantities for any other weight than one 
pound, it is necessary to multiply the values given in the table, 
except those of temperature and weight of 1 cu. ft. of steam, bj' 
the actual weight of the steam. 

The properties of steam depend on its absolute pressure. The 
pressure offers a certain amount of resistance to the expansion of 
the water into steam and it is the amount of this resistance which 
determines the temperature of evaporation and other quantities. 
Consequently, the absolute pressure is the first item given in the 
table. 

For convenience, the corresponding gage pressures are given 
in the next column, assuming an atmospheric pressure of 14.7 
lb. per sq. in. In case the barometer shows an atmospheric 
pressure very different from this, it is best to add the actual 
atmospheric pressure to the gage pressure and thus get the abso- 
lute pressure, which should then be used for finding the properties 
of the steam. In using properties of steam at pressures below 
that of the atmosphere, it is especially desirable to calculate the 
absolute pressure from barometer and vacuum gage readings 
rather than to use the vacuum gage reading in the gage pressure 
column of the table. An example will show what a difference 
this will make. 

Suppose a vacuum gage on a condenser shows a vacuum of 
27 inches and we want to find the temperature of the steam in the 
condenser. Without knowing the barometer reading but assum- 
ing the atmospheric pressure to be 14.7 lb. per sq. in., we would 
say that the reduction in pressure in the condenser amounted to 
27 X .4908 = 13.25 + lb. per sq. in. and that at a vacuum of 
13.25 lb. ( — 13.25 lb. gage), which corresponds to an absolute 
pressure of 14.7 — 13.25 = 1.45 lb. per sq. in., the temperature 
of the steam (from the steam table) is about 115°. 

Now suppose we first look at a barometer and find that it 
stands at 28 inches. The condenser has more vacuum than we 
thought it had. The absolute pressure in the condenser is 

(28 - 27) X .4908 = .4908 lb. per sq. in., 

or not quite .5 lb. per sq. in. absolute, and we find from the steam 
table that the temperature of the steam is a little less than 80° 
instead of being 115°. 

The third column in the table gives the temperatures at which 



PROPERTIES OF STEAM 67 

water boils when under the pressures given in the first column. 
These temperatures are also the temperatures of saturated steam 
under the given pressures, and likewise the temperatures at 
which steam under the given pressures will condense. 

The total heat required to form steam from water which has an 
initial temperature of 32° is found in the fifth column of the table. 
This quantity is the sum of the heat supplied to the water, and 
the latent heat. 

The heat of the liquid, or the heat in the water above 32°, is 
found in the third column. This is the amount of heat which 
must be supplied to the water to raise its temperature from 32° 
to the boiling point. For approximate calculations the heat of the 
liquid per pound of steam may be found by subtracting 32° from 
the boiling temperature, since the specific heat of water is nearly 
1, but for accurate calculations the heat of the liquid should be 
obtained from the steam table. The different results obtained 
by these two methods may be shown as follows: At 165 lb. per 
sq. in. absolute pressure the boiling temperature is 366°. This 
would give, by difference of temperature, 

366 - 32 = 334 B.t.u. 

for the heat of the liquid, while its actual value, from the steam 
table, is 338.2, a difference of over 1 per cent. 

After water is raised to the boiling point, heat must be added 
to change it into steam. This heat is called latent heat, and it 
varies in amount, decreasing as the pressure of the steam in- 
creases. At an absolute pressure of 14.7 lb. per sq. in., the latent 
heat is 970.4 B.t.u. per pound of steam, and at an absolute 
pressure of 100 lb. per sq. in. it is 888. B.t.u. The whole amount 
of the latent heat will be absorbed only when the whole pound 
of water has been evaporated; also, when one pound of steam is 
condensed, the full latent heat will be given up by it. If the 
water is being evaporated at 100 lb. per sq. in. absolute pressure 
and after reaching the boiling temperature only one-half of the 
latent heat or 

y^ X 888 = 444 B.t.u. 

are supplied to the water, then only one-half of a pound of steam 
will be formed, and conversely, if we extract 444 B.t.u. from a 
quantity of steam at 100 lb. per sq. in. absolute pressure, only 
one-half of a pound will be condensed. 

All of the quantities given in the steam table are calculated 



68 STEAM ENGINES 

from water at 32°, and in practical problems it is generally 
necessary to calculate the heat in steam above some other tem- 
perature than 32°. Thus if we wish to know how much heat must 
be supplied to one pound of water at 170° in order to turn it into 
steam having a pressure of 150 lb. per sq. in. absolute, we must 
remember that the water already contains 

170 - 32 = 138 B.t.u. 

Now, since the total heat of steam at 150 lb. per sq. in. absolute 
is 1193.4 B.t.u., there will have to be supplied only 

1193.4 - 138 = 1055.4 B. t. u. 

in order to turn it into steam. Since the heat of the liquid at 
150 lb. per sq. in. absolute pressure is 330.2 B.t.u., only 

330.2 - 138 = 192.2 B.t.u. 

need be supplied to the water to bring it to the boiling tempera- 
ture, but the entire latent heat, 863.2 B.t.u., must be supplied 
in order to evaporate it into steam. 

Interpolation from Tables.^ — Interpolation refers to the method 
used to find values between those given in the tables, as for 
example, finding the latent heat at 443-^ lb. per sq. in. absolute 
pressure. The table gives the latent heat for 44 lb. and for 45 
lb. but not for 44 J^ lb., and we interpolate to get the value for 
44J^ lb. which would be halfway between 929.2 and 928.2 or 
just 929.7. Suppose we wish to obtain the heat of the liquid 
at 120 lb. per sq. in. gage pressure (134.7 lb. absolute). The 
table gives 134 lb. and 135 lb., the corresponding values of the 
heat of the liquid being 321.1 and 321.7. For one pound change 
in pressure, the heat of the liquid changes 

321.7 - 321.1 = .6 B.t.u. 

Now, 134.7 is .7 lb. more than 134 or .3 lb. less than 135. We 
can, therefore, add .7 of .6 to 321.1 or subtract .3 of .6 from 321.7. 
Either way we get 321.52 as the value of the heat of the liquid 
at 120 lb. per sq. in. gage pressure. 

In interpolating, remember that the latent heat and the volume 
of one pound of steam decrease as the pressure increases and that 
all other items in the table increase. For most calculations it is 
sufficiently accurate to take the nearest value given in the table 
without bothering to interpolate. 



PROPERTIES OF STEAM 69 

Wet Steam. — Saturated steam may be either wet or dry. If 
wet, it has small particles of water suspended in it, just as in a 
fog air has particles of water suspended in it. The water which 
is suspended in wet steam has not been evaporated into steam and 
has not received the latent heat. It is in the form of water but 
is at boiling temperature and it has therefore received the entire 
heat of the liquid. Dry steam has no moisture in it and has 
received both the entire heat of the liquid and the entire latent 
heat. The quantities given in the steam tables are for dry 
saturated steam. 

In changing water into steam suppose that one pound of water 
is heated from 32° to the boiling temperature. Up to this point 
the one pound of water has received the entire heat of the liquid. 
When boiling commences, suppose that one-half of the pound of 
water is evaporated into steam and the other half of the pound is 
thrown up into the steam in the form of fine particles and remains 
suspended there. The steam has then received only one-half 
or .50 of the latent heat. If three-quarters of the water had been 
evaporated into the steam and the other quarter was suspended 
in the steam in the form of water, the steam would contain 
three-quarters or .75 of the latent heat. 

The total heat above 32° in one pound of dry steam is equal 
to the sum of the heat of the liquid and the latent heat, or calling 
the total heat H, the heat of the liquid h, and the latent heat L, 
then 

H = h+L 

The total heat in one pound of wet steam is the sum of the 
entire heat of the liquid, h, and that fraction of the latent heat 
which has formed steam, or 

H = h + qL 

in which q is the per cent, of the pound of steam which has been 
evaporated. The quantity, q, is called the quality of the steam. 
Example. — A pound of water at 32° is heated to the boiling temperature 
at a pressure of 100 lb. per sq. in. absolute and turned into steam having a 
quality of 90 per cent., that is, 90 per cent, of the pound of water is evapo- 
rated and the other 10 per cent, is suspended in the steam in the form of 
water. How much heat has been supplied to the steam? 

Solution. — For 100 lb. per sq. in. absolute 

h = 298.3 
L = 888.0 
6 



70 




STEAM ENGINES 


and in 


this case 


q = .9 


Therefore 


H = h + qL 






= 298.3 + .9 X 888.0 






= 298.3 + 799.2 






= 1097.5 B.t.u. 


Observ 


e that if the steam had been dry it would have cc 






H = h -hL 






= 298.3 + 888.0 






= 1186.3 B.t.u. 



Steam usually contains from 2 to 10 per cent, of moisture so that its 
quality is from 90 to 98 per cent. Steam which has a quality of 98 per cent, 
or more is called "commercially dry steam." 

In case the steam is formed from water having a temperature higher than 
32°, the heat which is already in the water above 32° must be subtracted 
from the total heat of the steam. For example, suppose steam at 150 lb. 
per sq. in. absolute pressure and having a quality of 95 per cent, is formed 
from water having a temperature of 170°. The heat already in the water 
above 32° is 

170 - 32 = 138 B.t.u. 

hence, only 330.2 — 138 = 192.2 B.t.u. need be supplied to the water per 
pound in order to bring it to the boiling temperature. Since the quality is 
95 per cent, only 

.95 X 863.2 = 820 B.t.u. 

is absorbed per pound in evaporating the water. Therefore, this steam 
has received 

192.2 + 820 = 1012.2 B.t.u. per pound 



Superheated Steam. — If saturated steam is taken away from 
the presence of water and heated, its temperature may be raised 
above that at which it was formed. Steam which has a higher 
temperature than that at which it was formed is called super- 
heated steam. Since superheated steam has received heat above 
that required to form it into saturated steam, it contains more 
heat per pound than saturated steam. 

The total heat above 32° contained in a pound of superheated 
steam may be found by adding to the total heat of saturated 
steam for the same pressure, as found in the steam table, the 
number of heat units required to superheat the steam, as shown 
by the following table: 



PROPERTIES OF STEAM 
Heat Units Required to Superheat Steam 



71 



Abso- 
lute 


Degrees of superheat 


pres- 
sure 


10 


20 


40 


60 


80 


100 


130 


160 


200 


250 


300 


1 


4.9 


9.6 


18.8 


27.9 


36.9 


46.0 


59.6 


73.2 


91.3 


114.0 


136.8 


10 


5.4 


10.4 


20.1 


29.6 


39.0 


48.4 


62.4 


76.3 


94.9 


118.0 


141.2 


15 


5.5 


10.6 


20.5 


30.2 


39.7 


49.2 


63.3 


77.4 


96.1 


119.4 


142.9 


20 


5.6 


10.8 


20.9 


30.7 


40.3 


49.9 


64.1 


78.3 


97.1 


120.6 


144.2 


30 


5.7 


11.1 


21.4 


31.4 


41.3 


51.0 


65.5 


79.8 


98.8 


122.6 


146.5 


40 


5.9 


11.3 


21.8 


32.0 


42.0 


51.9 


66.6 


81.1 


100.3 


124.2 


148.3 


50 


6.0 


11.5 


22.2 


32.5 


42.4 


52.6 


67.4 


82.1 


101.4 


125.6 


149.8 


60 


6.0 


11.7 


22.5 


32.9 


43.2 


53.3 


68.2 


82.9 


102.4 


126.7 


151.0 


80 


6.2 


11.9 


22.9 


33.6 


44.0 


54.2 


69.3 


84.2 


103.9 


128.4 


152.9 


100 


6.3 


12.2 


23.3 


34.1 


44.6 


55.0 


70.2 


85.2 


105.1 


129.7 


154.4 


130 


6.4 


12.4 


23.8 


34.7 


45.4 


55.8 


71.3 


86.4 


106.4 


131.2 


156.1 


160 


6.5 


12.6 


24.2 


35.3 


46.0 


56.6 


72.1 


87.4 


107.5 


132.5 


157.5 


200 


6.7 


12.9 


24.7 


35.9 


46.8 


57.4 


73.1 


88.6 


108.9 


134.1 


159.3 


250 


6.9 


13.2 


25.1 


36.5 


47.6 


58.4 


74.3 


89.9 


110.4 


135.9 


161.3 


300 


7.0 


13.5 


25.6 


37.1 


48.3 


59.2 


75.3 


91.0 


111.7 


137.4 


163.0 



The use of this table may be illustrated by the following example : 

Example. — Determine the number of heat units contained in a pound of 
superheated steam having a pressure of 130 lb. per sq. in. absolute and having 
a temperature of 447.4°. 

Solution. — By referring to the steam table we see that the temperature 
of saturated steam at a pressure of 130 lb. per sq. in. absolute is 347.4° and 
that its total heat is 1191 B.t.u. The degree of superheat is, therefore, 

447.4 - 347.4 = 100° 

And the above table shows that for this degree of superheat and for a 
pressure of 130 lb. per sq. in. absolute the number of heat units required 
to superheat the steam is 55.8. The pound of superheated steam will, 
therefore, contain 

1191 + 55.8 = 1246.8 B.t.u. 

Since superheated steam contains more heat than the same weight of satu- 
rated steam it is evident that the superheated steam is also dry. 



72 



STEAM ENGINES 



Properties of Dry Saturated Steam 
from Marks' and Davis' Steam Tables 



1 

Absolute 

pressure 

lb. per sq. in. 


2 

Tempera- 
ture; 
degrees 
Fahren- 
heit 


3 

Heat of 

the liquid 

per pound 

B.t.u. 


4 
Latent 
heat of 
evapora- 
tion per 
pound 
B.t.u. 


5 

Total 

heat per 

pound 

B.t.u. 


G 

Volume of 

one pound 

cu. ft. 


7 

Density or 

weight of 

one cu. ft. 

lbs. 


P 


t 


h 


L 


H 


V 


d 


0.0886 


32.0 


0.00 


1073.4 


1073.4 


3294.0 


0.000304 


1 


101.83 


69.8 


1034.6 


1104.4 


333.0 


0.00300 


2 


126.15 


94.0 


1021.0 


1115.0 


173.5 


0.00576 


3 


141.52 


109.4 


1012.3 


1121.6 


118.5 


0.00845 


4 


153.01 


120.9 


1005.7 


1126.5 


90.5 


0.01107 


5 


162.28 


130.1 


1000.3 


1130.5 


73.33 


0.01364 


6 


170.06 


137.9 


995.8 


1133.7 


61.89 


0.01616 


7 


176.85 


144.7 


991.8 


1136.5 


53.56 


0.01867 


8 


182.86 


150.8 


988.2 


1139.0 


47.27 


0.02115 


9 


188.27 


156.2 


985.0 


1141.1 


42.36 


0.02361 


10 


193.22 


161.1 


982.0 


1143. r 


38.38 


0.02606 


11 


197.75 


165.7 


979.2 


1144.9 


35.10 


0.02849 


12 


201.96 


169.9 


976.6 


1146.5 


32.36 


0.03090 


13 


205.87 


173.8 


974.2 


1148.0 


30.03 


0.03330 


14 


209.55 


177.5 


971.9 


1149.4 


28.02 


0.03569 


15 


213.0 


181.0 


969.7 


1150.7 


26.27 


0.03806 


16 


216.3 


184.4 


967.6 


1152.0 


24.79 


0.04042 


17 


219.4 


187.5 


965.6 


1153.1 


23.38 


0.04277 


18 


222.4 


190.5 


963.7 


1154.2 


22.16 


0.04512 


19 


225.2 


193.4 


961.8 


1155.2 


21.07 


0.04746 


20 


228.0 


196.1 


960.0 


1156.2 


20.08 


0.04980 


21 • 


230.6 


198.8 


958.3 


1157.1 


19.18 


0.05213 


22 


233.1 


201.3 


956.7 


1158.0 


18.37 


0.05445 


23 


235.5 


203.8 


955.1 


1158.8 


17.62 


0.05676 


24 


237.8 


206.1 


953.5 


1159.6 


16.93 


0.05907 


25 


240.1 


208.4 


952.0 


1160.4 


16.30 


0.0614 


26 


242.2 


210.6 


950.6 


1161.2 


15.72 


0.0636 


27 


244.4 


212.7 


949.2 


1161.9 


15.18 


0.0659 


28 


246.4 


214.8 


947.8 


1162.6 


14.67 


0.0682 


29 


248.4 


216.8 


946.4 


1163.2 


14.19 


0.0705 



PROPERTIES OF STEAM 



73 



Properties of Dry Saturated Steam — Continued 



Absolute 

pressure 

lb. per sq. in. 



Tempera- 
ture; 
degrees 
Fahren- 
heit 



Heat of 

the liquid 

per pound 

B.t.u. 



4 

Latent 
heat of 
evapora- 
tion per 
pound 
B.t.u. 



Total 

heat per 

pound 

B.t.u. 

H 



Volume of 

one pound 

cu. ft. 



Density or 

weight of 

one cu. ft. 

lbs. 



30 
31 
32 
33 
34 

35 
36 
37 
38 
39 

40 
41 
42 
43 
44 

45 

46 
47 
48 
49 

50 
51 
52 
53 
54 

55 
56 
57 

58 
59 

60 
61 
62 
63 
64 



250.3 


252.2 


254.1 


255.8 


257.6 


259.3 


261.0 


262.6 


264.2 


265.8 


267.3 


268.7 


270.2 


271.7 


273.1 


274.5 


275.8 


277.2 


278.5 


279.8 


281.0 


282.3 


283.5 


284.7 


285.9 


287.1 


288.2 


289.4 


290.5 


291.6 


292.7 


293.8 


294.9 


295.9 


297.0 ^ 



218.8 
220.7 
222.6 
224.4 
226.2 

227.9 
229.6 
231.3 
232.9 
234.5 

236.1 
237.6 
239.1 
240.5 
242.0 

243.4 
244.8 
246.1 

247.5 
248.8 

250.1 
251.4 
252.6 
253.9 
255.1 

256.3 
257.5 
258.7 
259.8 
261.0 

262.1 
263.2 
264.3 
265.4 
266.4 



945.1 1163.9 

943.8 1164.5 

942.5 1165.1 

941.3 1165.7 



940.1 

938.9 
937.7 
936.6 
935.5 
934.4 

933.3 
932.2 
931.2 
930.2 
929.2 

928.2 
927.2 
926.3 
925.3 
924.4 

923.5 
922.6 
921.7 
920.8 
919.9 

919.0 
918.2 
917.4 
916.5 
915.7 

914.9 
914.1 
913.3 
912.5 
911.8 



1166.3 

1166.8 
1167.3 
1167.8 
1168.4 
1168.9 

1169.4 
1169.8 
1170.3 
1170.7 
1171.2 

1171.6 
1172.0 
1172.4 
1172.8 
1173.2 

1173.6 
1174.0 
1174.3 
1174.7 
1175.0 

1175.4 
1175.7 
1176.0 
1176.4 
1176.7 

1177.0 
1177.3 
1177.6 
1177.9 
1178.2 



13.74 
13.32 
12.93 
12.57 
12.22 

11.89 
11.58 
11.29 
11.01 
10.74 

10.49 

10.25 

10.02 

9.80 

9.59 

9.39 
9.20 
9.02 

8.84 
8.67 

8.51 
8.35 
8.20 
8.05 
7.91 

7.78 
7.65 
7.52 
7.40 

7.28 

7.17 
7.06 
6.95 
6.85 
6.75 



0.0728 
0.0751 
0.0773 
0.0795 
0.0818 

0.0841 
0.0863 
0.0886 
0.0908 
0.0931 

. 0953 
0.0976 
. 0998 
0.1020 
0.1043 

0.1065 
0.1087 
0.1109 
0.1131 
0.1153 

0.1175 
0.1197 
0.1219 
0.1241 
0.1263 

0.1285 
0.1307 
0.1329 
0.1350 
0.1372 

0.1394 
0.1416 
0.1438 
0.1460 
0.1482 



74 



STEAM ENGINES 



Properties 


OF Dry Saturated Steam— 


-Continueo 


I 


1 

Absolute 

pressure 

lb. per sq. in. 


2 

Tempera- 
ture; 
degrees 
Fahren- 
heit 


3 

Heat of 

the liquid 

per pound 

B.t.u. 


4 
Latent 
heat of 
evapora- 
tion per 
pound 
B.t.u. 


5 

Total 

heat per 

pound 

B.t.u. 


6 

Volume of 

one pound 

cu. ft. 


7 

Density or 

weight of 

one cu. ft. 

lbs. 


P 


t 


h 


L 


H 


V 


d 


65 


298.0 


267.5 


911.0 


1178.5 


6.65 


0.1503 


66 


299.0 


268.5 


910.2 


1178.8 


6.56 


0.1525 


67 


300.0 


269.6 


909.5 


1179.0 


6.47 


0.1547 


68 


301.0 


270.6 


908.7 


1179.3 


6.38 


0.1569 


69 


302.0 


271.6 


908.0 


1179.6 


6.29 


0.1590 


70 


302.9 


272.6 


907.2 


1179.8 


6.20 


0.1612 


71 


303.9 


273.6 


906.5 


1180.1 


6.12 


0.1634 


72 


304.8 


274.5 


905.8 


1180.4 


6.04 


0.1656 


73 


305.8 


275.5 


905.1 


1180.6 


5.96 


0.1678 


74 


306.7 


276.5 


904.4 


1180.9 


5.89 


0.1699 


75 


307.6 


277.4 


903.7 


1181.1 


5.81 


0.1721 


76 


308.5 


278.3 


903.0 


1181.4 


5.74 


0.1743 


77 


309.4 


279.3 


902.3 


1181.6 


5.67 


0.1764 


78 


310.3 


280.2 


901.7 


1181.8 


5.60 


0.1786 


79 


311.2 


281.1 


901.0 


1182.1 


5.54 


0.1808 


80 


312.0 


282.0 


900.3 


1182.3 


5.47 


0.1829 


81 


312.9 


282.9 


899.7 


1182.5 


5.41 


0.1851 


82 


313.8 


283.8 


899.0 


1182.8 


5.34 


0.1873 


83 


314.6 


284.6 


898.4 


1183.0 


5.28 


0.1894 


84 


315.4 


285.5 


897.7 


1183.2 


5.22 


0.1915 


85 


316.3 


286.3 


897.1 


1183.4 


5.16 


0.1937 


86 


317.1 


287.2 


896.4 


1183.6 


5.10 


0.1959 


87 


317.9 


288.0 


895.8 


1183.8 


5.05 


0.1980 


88 


318.7 


288.9 


895.2 


1184.0 


5.00 


0.2001 


89 


319.5 


289.7 


894.6 


1184.2 


4.94 


0.2023 


90 


320.3 


290.5 


893.9 


1184.4 


4.89 


0.2044 


91 


321.1 


291.3 


893.3 


1184.6 


4.84 


. 2065 


92 


321.8 


292.1 


892.7 


1184.8 


4.79 


0.2087 


93 


322.6 


292.9 


892.1 


1185.0 


4.74 


0.2109 


94 


323.4 


293.7 


891.5 


1185.2 


4.69 


0.2130 


95 


324.1 


294.5 


890.9 


1185.4 


4.65 


0.2151 


96 


324.9 


295.3 


890.3 


1185.6 


4.60 


0.2172 


97 


325.6 


296.1 


889.7 


1185.8 


4.56 


0.2193 


98 


326.4 


296.8 


889.2 


1186.0 


4.51 


0.2215 


99 


327.1 


297.6 


888.6 


1186.2 


4.47 


0.2237 



PROPERTIES OF STEAM 



75 



Properties < 


DF Dry Saturated Steam — 


Continued 




1 

Absolute 

pressur 

lb. per sq. in. 


2 

Tempera- 
ture; 
degrees 
Fahren- 
heit 


3 

Heat of 

the liquid 

per pound 

B.t.u. 


4 
Latent 
heat of 
evapora- 
tion per 
pound 
B.t.u. 


5 

Total 

heat per 

pound 

B.t.u. 


6 

Volume of 

one pound 

cu. ft. 


7 

Density or 

weight of 

one cu. ft. 

lbs. 


P 


t 


h 


L 


H 


V 


d 


100 


327.8 


298.3 


888.0 


1186.3 


4.429 


0.2258 


101 


328.6 


299.1 


887.4 


1186.5 


4.388 


. 2279 


102 


329.3 


299.8 


886.9 


1186.7 


4.347 


0.2300 


103 


330.0 


300.6 


886.3 


1186.9 


4.307 


. 2322 


104 


330.7 


301.3 


885.8 


1187.0 


4.268 


0.2343 


105 


331.4 


302.0 


885.2 


1187.2 


4.230 


0.2365 


106 


332.0 


302.7 


884.7 


1187.4 


4.192 


0.2336 


107 


332.7 


303.4 


884.1 


1187.5 


4.155 


. 2408 


108 


333.4 


304.1 


883.6 


1187.7 


4.118 


. 2429 


109 


334.1 


304.8 


883.0 


1187.9 


4.082 


0.2450 


110 


334.8 


305.5 


882.5 


1188.0 


4.047 


0.2472 


111 


335.4 


306.2 


881.9 


1188.2 


4.012 


0.2593 


112 


336.1 


306.9 


881.4 


1188.4 


3.978 


0.2514 


113 


336.8 


307.6 


880.9 


1188.5 


3.945 


0.2535 


114 


337.4 


308.3 


880.4 


1188.7 


3.912 


0.2556 


115 


338.1 


309.0 


879.8 


1188.8 


3.880 


0.2577 


116 


338.7 


309.6 


879.3 


1189.0 


3.848 


. 2599 


117 


339.4 


310.3 


878.8 


1189.1 


3.817 


0.2620 


118 


340.0 


311.0 


878.3 


1189.3 


3.786 


0.2641 


119 


340.6 


311.6 


877.8 


1189.4 


3.756 


0.2662 


120 


341.3 


312.3 


877.2 


1189.6 


3.726 


. 2683 


121 


341.9 


313.0 


876.7 


1189.7 


3.697 


0.2705 


122 


342.5 


313.6 


876.2 


1189.8 


3.668 


0.2726 


123 


343.2 


314.3 


875.7 


1190.0 


3.639 


0.2748 


124 


343.8 


314.9 


875.2 


1190.1 


3.611 


0.2769 


125 


344.4 


315.5 


874.7 


1190.3 


3.583 


0.2791 


126 


345.0 


316.2 


874.2 


1190.4 


3.556 


0.2812 


127 


345.6 


316.8 


873.8 


1190.5 


3.530 


0.2833 


128 


346.2 


317.4 


873.3 


1190.7 


3.504 


0.2854 


129 


346.8 


318.0 


872.8 


1190.8 


3.478 


0.2875 


130 


347.4 


318.6 


872.3 


1191.0 


3.452 


0.2897 


131 


348.0 


319.3 


871.8 


1191.1 


3.427 


0.2918 


132 


348.5 


319.9 


871.3 


1191.2 


3.402 


0.2939 


133 


349.1 


320.5 


870.9 


1191.3 


3.378 


0.2960 


134 


349.7 


321.1 


870.4 


1191.5 


3.354 


0.2981 



76 



STEAM ENGINES 



Properties of Dry Saturated Steam — Continued 



Absolute 

pressure 

lb. per sq. in. 



Tempera- 
ture; 
degrees 
Fahren- 
heit 



Heat of 

the liquid 

per pound 

B.t.u. 



4 
Latent 
heat of 
evapora- 
tion per 
pound 
B.t.u. 



Total 

heat per 

pound 

B. .u. 

H 



6 

Volume of 

one pound 

cu. ft. 



Density or 

weight of 

one cu. ft. 

lbs. 



135 
136 
137 
138 
139 

140 
141 
142 
143 
144 

145 
146 
147 
148 
149 

150 
151 
152 
153 
154 

155 
156 
157 
158 
159 

160 
161 
162 
163 
164 

165 
166 
167 
168 
169 



350.3 
350.8 
351.4 
352.0 
352.5 

353.1 
353.6 
354.2 
354.7 
355.3 

355.8 
356.3 
356.9 
357.4 
357.9 

358.5 
359.0 
359.5 
360.0 
360.5 

361.0 
361.6 
362.1 
362.6 
363.1 

363.6 
364.1 
364.6 
365.1 
365.6 

366.0 
366.5 
367.0 
367.5 
368.0 



321.7 
322.3 
322.8 
323.4 
324.0 

324.6 
325.2 
325.8 
326.3 
326.9 

327.4 
328.0 
328.6 
329.1 
329.7 

330.2 
330.8 
331.4 
331.9 
332.4 

332.9 
333.5 
334.0 
334.6 
335.0 

335.6 

336.2 

336.7 

337.2' 

337.7 

338.2 
338.7 
339.2 
339.7 
340.2 



869.9 
869.4 
869.0 
868.5 
868.1 

867.6 
867.2 
866.7 
866.3 
865.8 

865.4 
864.9 
864.5 
864.0 
863.6 

863.2 
862.7 
862.3 
861.8 
861.4 

861.0 
860.6 
860.1 
859.7 
859.3 

858.8 
858.4 
858.0 
857.6 



857 


2 


856 


8 


856 


4 


855 


9 


855 


5 


855 


1 



1191.6 


3.331 


1191.7 


3.308 


1191.8 


3.285 


1192.0 


3.263 


1192.1 


3.241 


1192.2 


3.219 


1192.3 


3.197 


1192.5 


3.175 


1192.6 


3.154 


1192.7 


3.133 


1192.8 


3.112 


1192.9 


3.092 


1193.0 


3.072 


1193.2 


3.052 


1193.3 


3.033 


1193.4 


3.012 


1193.5 


2.993 


1193.6 


2.974 


1193.7 


2.956 


1193.8 


2.938 


1194.0 


2.920 


1194.1 


2.902 


1194.2 


2.885 


1194.3 


2.868 


1194.4 


2.851 


1194.5 


2.834 


1194.6 


2.818 


1194.7 


2.801 


1194.8 


2.785 


1194.9 


2.769 


1195.0 


2.753 


1195.1 


2.737 


1195.2 


2.72 


1195.3 


2.706 


1195.4 


2.690 



0.3002 
0.3023 
. 3044 
0.3065 
0.3086 

0.3107 
0.3129 
0.3150 
0.3171 
0.3192 

0.3213 
0.3234 
0.3255 
0.3276 
0.3297 

0.3320 
0.3341 
0.3362 
0.3383 
0.3404 

0.3425 
0.3446 
0.3467 
0.3488 
0.3508 

0.3529 
0.3549 
0.3570 
0.3591 
0.3612 

0.3633 
0.3654 
0.3675 
. 3696 
0.3717 



PROPERTIES OF STEAM 



77 



Properties of Dry Saturated Steam — Continued 



1 

Absolute 

pressure 

lb. per sq. in. 


2 

Tempera- 
ture; 
degrees 
Fahren- 
heit 


3 

Heat of 

the liquid 

per pound 

B.t.u. 


4 
Latent 
heat of 
evapora- 
tion per 
pound 
B.t.u. 


5 

Total 

heat per 

pound 

B.t.u. 


6 

Volume of 

one pound 

cu. ft. 


7 

Density or 

weight of 

one cu. ft. 

lbs. 


P 


t 


h 


L 


H 


V 


d 


170 


368.5 


340.7 


854.7 


1195.4 


2.675 


. 3738 


171 


368.9 


341.2 


854.3 


1195.5 


2.660 


0.3759 


172 


369.4 


341.7 


853.9 


1195.6 


2.645 


0.3780 


173 


369.9 


342.2 


853.5 


1195.7 


2.631 


0.3801 


174 


370.4 


342.7 


853.1 


1195.8 


2.616 


. 3822 


175 


370.8 


343.2 


852.7 


1195.9 


2.602 


0.3843 


176 


371.3 


343.7 


852.3 


1196.0 


2.588 


0.3864 


177 


371.7 


344.2 


851.9 


1196.1 


2.574 


0.3885 


178 


372.2 


344.7 


851.5 


1196.2 


2.560 


0.3906 


179 


372.7 


345.2 


851.2 


1196.3 


2.547 


0.3927 


180 


373.1 


345.6 


850.8 


1196.4 


2.533 


0.3948 


181 


373.6 


346.1 


850.4 


1196.5 


2.520 


0.3969 


182 


374.0 


346.6 


850.0 


1196.6 


2.507 


0.3989 


183 


374.5 


347.1 


849.6 


1196.7 


2.494 


0.4010 


184 


374.9 


347.6 


849.2 


1196.8 


2.481 


0.4031 


185 


375.4 


348.0 


848.8 


1196.8 


2.468 


0.4052 


186 


375.8 


348.5 


848.4 


1196.9 


2.455 


0.4073 


187 


376.3 


349.0 


848.0 


1197.0 


2.443 


0.4094 


188 


376.7 


349.4 


847.7 


1197.1 


2.430 


0.4115 


189 


377.2 


349.9 


847.3 


1197.2 


2.418 


0.4136 


190 


377.6 


350.4 


846.9 


1197.3 


2.406 


0.4157 


191 


378.0 


350.8 


846.5 


1197.3 


2.393 


0.4178 


192 


378.5 


351.3 


846.1 


1197.4 


2.381 


0.4199 


193 


378.9 


351.7 


845.8 


1197.5 


2.369 


0.4220 


194 


379.3 


352.2 


845.4 


1197.6 


2.358 


0.4241 


195 


379.8 


352.7 


845.0 


1197.7 


2.346 


0.4262 


196 


380.2 


353.1 


844.7 


1197.8 


2.335 


0.4283 


197 


380.6 


353.6 


844.3 


1197.8 


2.323 


0.4304 


198 


381.0 


354.0 


843.9 


1197.9 


2.312 


0.4325 


199 


381.4 


354.4 


843.6 


1198.0 


2.301 


0.4346 


200 


381.9 


354.9 


843.2 


1198.1 


2.290 


0.437 


205 


384.0 


357.1 


841.4 


1198.5 


2.237 


0.447 


210 


386.0 


359.2 


839.6 


1198.8 


2.187 


0.457 


215 


388.0 


361.4 


837.9 


1199.2 


2.138 


0.468 


220 


389.9 


363.4 


836.2 


1199.6 ' 


2.091 


0.478 



78 



STEAM ENGINES 



Properties < 


3F Dry Saturated 


Steam — 


Continued 




I 

Absolute 

pressure 

lb. per sq. in. 


2 

Tempera- 
ture; 
degrees 
Fahren- 
heit 


3 

Heat of 

the liquid 

per pound 

B.t.u. 


4 
Latent 
heat of 
evapora- 
tion per 
pound 
B.t.u. 


5 

Total 

heat per 

pound 

B.t.u. 


6 

Volume of 

one pound 

cu. ft. 


7 . 

Density or 

weight of 

one cu. ft. 

lbs. 


P 


t 


h 


L 


H 


V 


d 


225 


391.9 


365.5 


834.4 


1199.9 


2.046 


0.489 


230 


393.8 


367.5 


832.8 


1200.2 


2.004 


0.499 


235 


395.6 


369.4 


831.1 


1200.6 


1.964 


0.509 


240 


397.4 


371.4 


829.5 


1200.9 


1.924 


0.520 


245 


399.3 


373.3 


827.9 


1201.2 


1.887 


0.530 


250 


401.1 


375.2 


826.3 


1201.5 


1.850 


0.541 


260 


404.5 


378.9 


823.1 


1202.1 


1.782 


0.561 


270 


407.9 


382.5 


820.1 


1202.6 


1.718 


0.582 


280 


411.2 


386.0 


817.1 


1203.1 


1.658 


0.603 


290 


414.4 


389.4 


814.2 


1203.6 


1.602 


0.624 


300 


417.5 


392.7 


811.3 


1204.1 


1.551 


0.645 



CHAPTER VI 



INDICATORS 



Work Diagrams. — A diagram may be drawn, by means of an 
instrument called an indicator, which shows the work performed 
by the steam in the cylinder of a steam engine, and from such a 
diagram may be calculated the horsepower developed by the 
engine. 

The area of any diagram is equal to the product obtained by 
multiplying together its two sides. For example, in Fig. 49 
the figure ahcd has one side equal to 6 feet and the other side equal 
to 3 feet, hence the area of the diagram ahcd is 6 X 3 = 18 
sq. ft. In Fig. 50 is a similar diagram, except that one side repre- 



'5 


a 


b 


4 








■^ 3 


■ 




c 






o 


1 


- 


. 





3 4- 

Fig. 49. 



sents distance in feet and the other side represents force in pounds. 
The area of this diagram ahcd is equal to the product of the two 
sides or 6 X 3 = 18, but in this case the area represents foot- 
pounds, since the product of feet and pounds is foot-pounds or 

6 X 3 = 18 foot-pounds. 

Since foot-pounds is the unit of work, the area of the diagram 
ahcd in Fig. 50, represents work, or is a work diagram. In a 
8 79 



80 



STEAM ENGINES 



similar manner, if a diagram is drawn so that one side represents 
distance and the other side represents force (or pressure, which is 
a force) the area of the diagram will represent foot-pounds of 
work. This is the same principle upon which an indicator draws 
a work diagram for the cylinder of a steam engine. The indicator 
is so arranged that it draws on a sheet of paper a diagram which 
represents by its height the pressure of the steam in the cylinder 
and by its length represents the stroke of the piston. Such a 
diagram, with one side representing distance and the other side 
representing pressure, shows by its area the work being performed 
in the cylinder. 




The Indicator. — An indicator must record the pressure in the 
cylinder at every part of the stroke. In order to do this, the 
indicator consists of two parts, one of which moves in a vertical 
direction, with the steam pressure in the cylinder and the other 
which moves in a horizontal direction in unison with the piston. 
The part moving horizontally in unison with the piston carries a 
sheet of paper and the part moving vertically in unison with the 
pressure carries a pencil point. When the pencil point is brought 
into contact with the paper, a diagram is drawn which shows the 
pressure in the cylinder at every part of the stroke. 

One type of indicator, shown in Fig. 51, is partly cut away 
in order to show the inside construction of it. The device for 
measuring the steam pressure consists of a small cylinder 5, 
fitted with a piston 8, the cylinder being attached to the clearance 
space of the engine cylinder so that the steam pressure in the 



INDICATORS 



81 



clearance space acts upon the under side of the indicator piston. 
A piston rod 10 is connected to the piston by means of a ball 
joint in order to give flexibihty between the rod and the piston. 
The piston rod passes through the cyhnder head 2, and its end is 
joined to the pencil arm 16 by means of a short link 14. One end 
of the pencil arm is pivoted at 18. The other end, which moves 
over a sheet of paper on the drum 24, carries a pencil point, and 
is forced to move in a vertical hne by the Unks 14 and 15. The 
piston is held down by a finely adjusted coil spring and the 
steam pressure forces the piston upward against the resistance 




Fig. 51. 

of this spring. These springs are interchangeable and several 
are supplied with each indicator. The number of the spring 
indicates the number of pounds pressure per square inch which, 
acting upon the piston, will move the pencil point one inch; for 
example, if a No. 60 spring is in the indicator it will require a 
pressure of 60 lb. per sq. in. on the piston to move the pencil 
point one inch vertically. Therefore, the diagram drawn by the 
indicator may be measured and the pressure in the engine cylinder 
at any point of the stroke determined. An indicator spring 
should always be used with the instrument for which it is intended 
as its scale depends upon the length of the pencil arm and a slight 
difference in the length of the arm makes considerable difference 



82 STEAM ENGINES 

in the movement of the pencil, since the movement of the pencil 
point is usually five times the movement of the piston. The 
pencil arm, together with its links, is free to turn about the axis 
of the cylinder so the pencil point may be pressed against the drum 
or lifted away from it. 

The piston is ground to a close fit with the cylinder in order 
to make a nearly steam tight fit and, at the same time, not cause 
excessive friction. Any steam that leaks past the piston escapes 
through the hole A in the cylinder and is thus prevented from 
collecting over the piston and exerting a downward pressure upon 
it. Phis piston has a number of small grooves cut in its edge, 
which serve to hold lubricating oil and also aid in preventing 
the leakage of steam. 

The drum 24, which carries the sheet of paper and which moves 
in unison with the engine piston, is placed parallel to the indicator 
cylinder and is located at the end of the arm L, which forms a 
part of the indicator cylinder. The drum is clamped to the end 
of the arm L, by means of the thumb nut 39. A cord wrapped 
around the base of the drum in the groove 27 serves to turn it 
about its vertical axis. The cord takes its motion from the 
crosshead, which has the same motion as the piston. The cord 
is not usually attached directly to the crosshead because the 
stroke of the crosshead is greater than the circumference of the 
drum and would therefore turn it through more than one complete 
revolution, which would be undesirable on account of the pencil 
point striking the paper clips, 40. The drum is usually provided 
with stops which prevent it from turning through a complete 
revolution. 

The circumference of the drum is about 5 inches and it is de- 
sirable to have the indicator diagram about 3 to S}i inches long, 
while the crosshead may have a stroke of 2 or 3 feet, hence it is 
necessary to use some device which copies the motion of the 
crosshead on a reduced scale and to attach the cord to this device, 
which is called a reducing motion. Various forms of reducing 
motions will be described later. 

On the forward stroke of the crosshead the cord which is 
wrapped around the drum is pulled outward, turning the drum 
through a part of a revolution. At the same time a coil spring 
inside the drum is wound up. This coil spring has one end 
attached to the drum and the other end to the stationary 
spindle 28, hence, when the crosshead makes the return stroke the 



INDICATORS 83 

drum turns in the opposite direction, keeping the cord taut and 
rewinding it in its groove 27. 

All the moving parts of an indicator are made as light as 
possible, to avoid inertia effects, or over-travelling. This is 
especially necessary with the piston and connected parts, as 
these move rapidly, and inertia would cause the pencil point 
to move too high on its upward stroke and too low on its down- 
ward stroke. For the same reason, it is advisable to have a 
rather stiff spring in the indicator, or one which will give a dia- 
gram about 1}^ to 2 inches high. The number of the spring to 
use depends both upon the boiler pressure and the speed of the 
engine; thus with a boiler pressure of 90 lb. per sq. in. a No. 60 
spring should be used with a high speed engine, as this will 
give a diagram with a maximum height of ^%o = 1/^ inches. 
The number of spring to be used with any pressure may be found 
by dividing the pressure by the desired height of diagram, but it 
should be remembered that a high speed engine requires a 
stiff er spring than a slow speed one. 

Example. — "What number of indicator spring should be used with a boiler 
pressure of 125 lb. per sq. in. if it is desired to obtain a diagram 13^ inches 
high? 

Solution. — 7Y7 = 83.3 

i-72 

As the nearest regular size of spring is 80, this would be used. The regular 
sizes of springs are 8, 10, 12, 16, 20, 30, 40, 50, 60, 80, 100, 120, 150, and 180. 

An indicator is attached to a cylinder by means of a short 
length of pipe and a quick-opening valve placed just below it. 
This valve, which is furnished with the instrument, is used so 
that steam pressure may be cut off when the indicator is not being 
used, thus reducing wear on the working parts. 

All engine cylinders have holes at each end which are bored and 
threaded for attaching indicators, the holes entering the clearance 
space so they will not be covered by the piston at any time. If 
possible, it is best to use two indicators, one attached to each 
end of the cylinder, as this makes it possible to take diagrams 
from the two ends of the cylinder at the same time and also per- 
mits shorter connections between the indicator and cylinder. A 
long indicator connection is likely to cause a drop in pressure 
and thus make the indicator give a false record. Straightway 
valves are used for the same reason. 



84 



STEAM ENGINES 



A single indicator is sometimes attached as shown in Fig. 52 
and used for taking diagrams from both ends of the cyhnder, in 
which case both diagrams are drawn on the same sheet of paper. 
The principal disadvantage in using a single indicator is the time 
required to shut off steam from one end of the cylinder and turn 




Fig. 52. 



it on from the other end, which does not allow diagrams to be 
taken from the two ends of the cylinder at the same time. The 
time elapsing between taking the diagrams may, however, be 
greatly shortened by using a three-way valve, as shown in Fig. 




Fig. 53. 

53, at the indicator, instead of the straightway valve shown in 
Fig. 52. 

Indicators made by the various manufacturers differ from each 
other in small details, but most of them have the general form 
illustrated by Fig. 51. An indicator having a different form of 



INDICATORS 



85 



pencil movement is shown in Fig. 54. In this indicator one of 
the usual movable hnks is replaced by a slot cut in a plate G 
and the pencil arm has a small roller on its side which moves in 
this slot. The slot is for the purpose of giving a perfect straight 
Hne motion to the pencil point and for securing a uniform pro- 
portion between the motions of the pencil point and indicator 
piston, and it is shaped to secure these results. In this indicator, 




Fig. 54. 



the drum spring is a flat coil spring placed at the base of the drum, 
as shown at M. 

Fig. 55 shows a form of indicator having the spring outside the 
cylinder, the piston rod being made longer to hold it. When the 
spring is placed inside the cylinder and used with high pressure 
steam, the high temperature to which it is subjected is liable to 
change its stiffness and hence its scale. The outside spring ar- 
rangement is intended to overcome this disadvantage as well as 
to simplify the operation of changing springs for different pres- 



86 STEAM ENGINES 

sures. The outside spring is especially adapted to superheated 

steam. 

The indicator shown in Fig. 55 also illustrates another recent 
improvement in indicators, that is, a drum with which a number 
of diagrams may be drawn on the same paper. It is designed to 
use a roll of paper upon which the indicator traces a series of 
diagrams which will continue until the roll is exhausted, unless 




Fig. 55. 

interrupted by the operator. The roll of paper is located within 
an opening in the shell of the drum, thence the paper passes 
around the outside of the drum and inward to a central cyUnder, 
to which it is attached. Upon the top of the drum is a ratchet 
wheel which automatically unwinds a small length of paper from 
the spool and winds it on the inner cyhnder, thus giving a series 
of diagrams which overlap each other slightly, as shown in Fig. 
56. These indicators are commonly used with engines in which 
the load changes rapidly, such as rolHng mill engines, because they 



INDICATORS 



87 



c 



record the changes in load and also show the action of the 
under such changes. The other ordinary form 
of indicator is suitable for drawing a diagram 
during only a single revolution, and if the pencil 
point is held on the paper for more than a single 
revolution, the diagrams will be drawn upon each 
other, making it difficult to distinguish them. 

Reducing Motions. — The principal require- 
ment of a reducing motion is that it shall re- 
produce accurately the motion of the crosshead 
on a small scale. Some reducing motions which 
are very simple in construction do not reproduce 
the motion of the crosshead accurately. 

A form of reducing motion intended to be at- 
tached directly to the indicator is illustrated in 
Fig. 56a. This reducing motion consists of a 
large pulley around which is wrapped the cord 
which connects with the crosshead of the engine. 
This pulley drives a smaller pulley by means of 
bevel gears which reduce the motion. The smaller 
pulley drives the indicator drum by means of a 
cord wrapped around the drum and the smaller 
pulley. This reducing motion is supplied with 
several different sizes of the smaller pulley which 
adapts it to steam engines having piston strokes 
ranging from 14 inches to 72 inches. ■ 

This type of reducing motion will reproduce 
the motion of the crosshead accurately if cords 
are used which do not stretch and if the cords 
are prevented from piling on top of each other as 
they wrap around the pulleys. 

The reducing motion shown in Fig. 52 is one of 
the simplest and most common forms. It consists 
of a wooden arm pivoted at the top to a stand 
which is attached either to the floor or to the frame 
of the engine. The lower end of this arm has a 
slot for receiving a pin fitted to the center of the 
wrist pin. The upper end of the arm has a 
curved block fastened to it with a groove in its 
edge in which is placed the cord to the indicator. 
In order to reproduce accurately the motion of 



engme 









J 



88 



STEAM ENGINES 



the crosshead the curvature of the block must be such that the 
distance ob from the pivot to the center of the wrist pin divided 
by the distance oa from the pivot to the center of the cord must 




Fig. 56a. 



be constant at all points of the stroke, and also the pivot 
must be directly over the center of the wrist pin when the 




Fig. 57. 



crosshead is at the middle of its stroke. Lost motion between 
the slot and the pin in the wrist pin must also be avoided. 
The top view of the engine illustrated in Fig. 57, shows a 



INDICATORS 89 

pantograph reducing motion; one end of the pantograph being 
attached to the crosshead and the other end to a stand placed on 
the floor. The indicator cord is attached to the crossbar a, 
which is attached to the bars b and c. The reducing motion 
may be used with different lengths of stroke by attaching the 
crossbar a at various points along the bars b and c, but always 
fastening it in corresponding holes in b and c. On account of 
the large number of joints in the pantograph and the probability 
of lost motion in them, this form of reducing motion should be 
made of steel with closely fitting joints. When so made the 
pantograph reproduces the motion of the crosshead accurately. 
In attaching the cord to the reducing motion, it should always 




POINT OF 

XHAUST OPENINQ 



POINT OF EXHAUST CLOSURE: ATM05/=>H£f?/C /^/Rf-S^^/P^ Z/zV/ET 

Fig. 58. 

be arranged so it will run in a direction parallel to the crosshead, 
as in Fig. 57, otherwise the motion which the drum receives will 
not be the same as that of the crosshead. Also the cord must be 
attached to cross bar a on the center line of the pantograph or a 
correct reduction of the crosshead movement will not be obtained. 
Several holes should therefore be provided in cross bar a for 
the insertion of the pin so that the latter may be properly located 
as the bar a is moved from one position to another on bars b and 
c for different lengths of stroke. 

Indicator Diagrams. — In taking an indicator diagram, a paper 
card, especially prepared for the purpose, is placed on the drum 
and the cord attached to the reducing motion. Steam is then 
turned into the indicator and after it is warmed up the pencil 
point is touched to the drum while the engine is making a single 
revolution. Steam is then turned off the indicator and the pencil 
point again touched to the drum in order to draw the atmospheric 
line. The resulting diagram will be similar to that shown in Fig. 58. 



90 STEAM ENGINES 

Since the atmospheric Hne is drawn while only the pressure of 
the atmosphere is acting upon the piston of the indicator, this 
line represents the pressure of the atmosphere and it serves as a 
reference line for other parts of the diagram. Gage pressures 
may be measured from the atmospheric line, but if absolute 
pressures are desired, it will be necessary to draw a line of no 
pressure parallel with the atmospheric line and at a distance below 
it equal to the atmospheric pressure as read on a barometer, and 
drawn to the same scale to which the diagram is drawn. 

The indicator diagram shows the varying pressure in the 
cylinder for a complete revolution or a forward and back stroke, 
and anything which affects this pressure also affects the shape of 
the diagram. It also shows, by its area, the work being per- 
formed in the cylinder. The method of calculating the work from 
the diagram will be given later. 

The diagram shown in Fig. 58 is from only one end of the cylin- 
der, the end towards the left, but since it shows all changes of 
pressure, it gives a record of a complete cycle of the events 
taking place in that end of the cylinder. These events have 
been marked on the diagram for reference. The point at which 
steam is admitted to the cylinder is shown at the left, slightly 
before the piston reaches the end of its back stroke. As soon as 
steam is admitted the pressure in this end of the cylinder rises 
to the full admission pressure. By the time this has occurred 
the piston has reached the end of its return stroke and is starting 
on its forward stroke. During the first part of the forward stroke 
steam is being admitted behind the piston, hence the pressure 
remains constant during this part of the stroke. The part of 
the diagram drawn while steam is being admitted is called the 
steam line. At the end of the steam line is the point of cut-off 
at which the admission valve closes. On account of the gradual 
closing of this valve, the pressure changes gradually and the point 
of cut-off is not sharply defined. After the admission valve 
closes, the steam in the cylinder expands behind the advancing 
piston, as shown by the expansion line, the pressure of the steam 
gradually becoming smaller as its volume increases. Just before 
the piston has completed its forward stroke the exhaust valve 
opens and the pressure of the steam quickly drops while the piston 
is completing its stroke. The exhaust valve remains open dur- 
ing the greater part of the return stroke, and the piston pushes 
the low pressure steam from the cyhnder, giving the exhaust Hne 



INDICATORS 



91 



on the diagram. As the ports and exhaust passages offer a cer- 
tain amount of resistance to the flow of the exhaust steam, the 
exhaust Hne will be above the atmospheric line by a distance 
which represents a few pounds pressure. This pressure is called 
the '^ back-pressure" since it acts against the advancing piston. 




Fig. 59. 

Near the end of the exhaust stroke the exhaust valve closes, 
giving the '^ point of compression" or ''exhaust closure." After 
the exhaust valve is closed the steam remaining in the cylinder 
is compressed until the admission valve opens, thus completing 
the cycle of events for this end of the cylinder. 

Similar events occur in the other end of the cylinder but not at 
the same time. These may be shown on a separate diagram 




Fig. 60. 

drawn with another indicator, or, if a single indicator is used for 
both ends of the cyHnder, both diagrams will be drawn on one 
paper or ''card," and this will show the relative positions of the 
events. Such a double diagram is illustrated in Fig. 59, which 
shows that admission and expansion are occurring in the head 
end of the cylinder while exhaust is taking place from the crank 
end, and that admission and expansion are occurring in the crank 
end while exhaust is taking place from the head end. 



92 



STEAM ENGINES 



Besides being used to determine the power developed by an 
engine, the indicator diagram also shows whether or not the 
engine and indicator are adjusted properly. Faults in the engine 
adjustment will be considered in a later chapter. A few of the 
more common indicator faults will be considered here. Fig. 60 
shows diagrams taken by an indicator in which the cord is too 
long, thus allowing the drum to stop before the crosshead has 




Fig. 61. 

completed its stroke. It will be seen that the left-hand ends of 
both diagrams appear to be cut off, the heel of one diagram and 
the toe of the other being cut off on the same Une. The same 
fault may be caused by the motion of the crosshead not being 
reduced sufficiently, but in this case the cord is liable to be broken. 
Sometimes the piston of a new indicator will fit too tightly, 
causing it to stick in the cyhnder. The result will show in a 




Fig. 62. 

stepped expansion Une as in Fig. 61. The steps will usually be 
more distinct near the beginning of the expansion line where the 
pressure is high. The same fault may be caused in an old indi- 
cator by a gummed piston which has not been cleaned and 
lubricated. 

If the spring used in an indicator is too weak for the pressure, 
the diagram will not only be too high, but its steam Une wiU be 



INDICATORS 93 

wavy, especially near the end, as shown in Fig. 62. Such a wavy 
line is caused by the vibration of the spring when high pressure 
steam is first admitted to the cylinder. The remedy for this is, 
of course, to use a stiffer spring. 

Expansion of Steam. — Between the point of cut-off and release 
the weight of steam in the cylinder remains constant provided 
there is no leakage of steam either into or out of the cylinder. 
The steam that is in the cylinder simply expands, that is, its 
volume increases and its pressure falls. 

If we should select a number of points along the expansion line 
of an indicator diagram and multiply the absolute pressure at 
each of these points by the volume of steam in the cylinder at 
that point we would find that the product of this multiplication 
would be practically a constant number. This being true it is 
evident that the pressure of the steam falls at the same rate that 
its volume increases. When the volume of the steam has in- 
creased to twice the volume contained in the cylinder at the point 
of cut-off, the absolute pressure of the steam will be one-half of 
what it was at the point of cut-off. In like manner, when the 
steam has expanded so that its volume is 1.5 times its volume at 

cut-off its absolute pressure will be ^ or r _ of what it was at 

the point of cut-off; and when the steam has expanded so that 
its volume is 4 times its volume at cut-off, its absolute pressure 
will be 3^^ of its absolute pressure at cut-off. It should be 
observed that the volume of steam which is expanding refers 
to the total volume of steam which is in the cylinder when cut-off 
occurs. This volume includes not only the volume of steam 
taken into the cylinder at each stroke, which is the same as the 
volume displaced by the piston from the beginning of its stroke 
up to the point of cut-off, but it includes also the volume of 
steam in the clearance space when the piston is at the beginning 
of its stroke. 

Example. — Cut-off occurs at ^g stroke in a 10" X 12" engine having 12 
per cent, clearance. What is the total volume of steam in the cyhnder at 
the beginning of expansion? 

Solution. — The area of the piston is 

^^j^^ X .7854 = .5454 sq. ft. 

Since the length of stroke is 12 in. or 1 ft., the piston displacement is 

.5454 X 1 = .5454 cu. ft. 



94 



STEAM ENGINES 



and the clearance volume is 

.5454 X .12 = .0854 cu. ft. 
Since cut-off occurs at % stroke the volume of steam taken into the cylinder 
at each stroke is 

% X .5454 = .3409 cu. ft. 
and the total volume of steam at the beginning of expansion is 

.3409 + .0654 = .4033 cu. ft. 

The quantities calculated in the above example are shown on 
the indicator diagram illustrated in Fig. 63. In this illustration 
the piston displacement is represented by the length of the dia- 
gram. The clearance volume c which in this case, is 12 per cent, 
of the piston displacement is represented by the distance between 
the end of the diagram and the line ao, which is the line of no 




Fig. 63. 



volume. The line ao is located by making the distance c equal 
to the clearance volume to the same scale that the length of the 
indicator card represents the piston displacement. For example, 
suppose the indicator diagram is 2.5 in. long, and this 2.5 in. 
represents the piston displacement of the above example or .5454 

2.5 

In this case a length of 



cu. 



ft. 



or 4.583 inches would 



5454 

represent 1 cu. ft. of volume. Therefore the clearance volume c, 
which is .0654 cu. ft. would be represented by a length of .0654 X 
4.583 = .2997 in. or practically .3 in. That is, the no volume 
line ao would be drawn .3 in. from the end of the indicator 
diagram. 

Ratio of Expansion. — The ratio of expansion is a measure of the 
number of times the steam is expanded in the cylinder. For 
example, if there are two cubic feet of steam in the cylinder when 



INDICATORS 95 

cut-off occurs (at the beginning of expansion) and this is expanded 
to four cubic feet, its volume has been increased two times, or its 
ratio of expansion is two. . 

Since the increase in volume of steam during expansion is 
practically in proportion to the decrease in pressure, the pressure 
of the steam at the end of expansion may be calculated, if the 
admission pressure and the ratio of expansion are known. 

Example. — If the admission pressure is 60 lb. per sq. in. and the ratio 
of expansion is 4, what will be the pressure in the cylinder at the end of 
expansion ? 

Solution. — Pressure at end of expansion 

= ^ = 15 lb. per sq. m. 

It may be seen from the above discussion and example that for 
a given admission pressure the final pressure will be lowest with 
an early cut-off, or large ratio of expansion. Also, if cut-off 
occurs late in the stroke, the steam being expanded but little, 
the pressure will be high at the end of the stroke when the exhaust 
valve opens. In the latter case, the pressure remaining in the 
steam is wasted, hence an early cut-off or large ratio of expansion 
is more desirable than a small ratio of expansion. 

The number of times which steam may be expanded in a 
cylinder depends upon the admission pressure of the steam as well 
as upon the point of cut-off. If this pressure is low and the 
steam is expanded a large number of times, the final pressure 
will be carried below the exhaust pressure, forming a loop at the 
toe of the diagram, and no useful work will be gained from the 
last part of the expansion. For example, if the admission pressure 
is 60 lb. per sq. in. and the ratio of expansion is 6, the final pres- 
sure of the steam will be 10 lb. per sq. in., which is below atmos- 
pheric pressure. If this engine exhausts into the atmosphere, 
the expansion below 14.7 lb. per sq. in. produces no useful work 
because, when the exhaust valve opens, the pressure in the cylin- 
der will rise to 14.7 lb. per sq. in. In fact, the pressure in the 
cylinder at the end of expansion should be a few pounds above 
the exhaust pressure because, if the steam is expanded completely 
to the exhaust pressure, the extra work gained is not enough to 
compensate for the friction of the engine during the last part of 
the stroke; hence, instead of there being a gain from the expan- 
sion of the last few pounds of pressure, there is actually a loss. 

An approximate value of the ratio of expansion may be taken 



96 



STEAM ENGINES 



as the reciprocal of the fraction of the piston stroke at which cut- 
off occurs. For example, if cut-off occurs at }^ stroke this 
approximate value of the ratio of expansion is 4; if cut-off oc- 
curs at % stroke the approximate ratio of expansion is ^^ or 13^:3. 
This method of computing the ratio of expansion gives only an 
approximate value because the clearance volume is neglected. 

Whether the ratio of expansion is calculated by the exact 
method or the approximate method it is necessary to locate the 
point of cut-off on the indicator diagram. As the point of cut- 
off is not sharply defined on the indicator diagram it is rather 
difficult to locate the exact point of cut-off; but it may be done 
with a fair degree of accuracy by locating the point of cut-off 
at the point where the downward curve of the admission line 
meets the upward curve of the expansion line. 




Fig. 64. 

The difficulty of locating the point of cut-off exactly on indica- 
tor diagrams has led to the use of the commercial cut-off in deter- 
mining the ratio of expansion. The commercial cut-off is 
located by drawing a horizontal line on the diagram through the 
maximum admission pressure and extending the expansion line 
up to meet this line. The intersection of these two lines is the 
commercial cut-off. 

The method of determining the commercial cut-off and from 
it the ratio of expansion is illustrated in Figs. 64 and 65. The 
line EH is first drawn so that the length EC represents the 
clearance volume to the same scale that the diagram is drawn. 
The line EA is then drawn through the maximum admissiom 
pressure and parallel to the atmospheric line. In case the 
admission line is wavy, as in Fig. 65, the line EA is drawn at the 
average height of the waves. From the point D where the ex- 



INDICATORS 



97 



pansion line changes direction of curvature the expansion Hne 
is extended upward to intersect the Une EA at the point B. 
The point B is the point of commercial cut-off. The fraction of 
stroke up to the commercial cut-off is 

BC 
AC 




H G 



Fig. 65. 



and the ratio of expansion is 

AC -\-EC 
BC + EC 



AE 
BE 



These distances may be measured directly on the diagram or 
they may be calculated from the piston displacement and clear- 
ance volume, if the point of commercial cut-off is known. 



CHAPTER VII 

INDICATED AND BRAKE HORSEPOWER 

Mean Effective Pressure. — The area of the diagram, which 
represents the work being performed in the cyhnder, may be 
found by multiplying together its height and length, having 
proper regard for the scales of pressure and stroke, but since the 
diagram is of irregular shape its average height must be used in 
this calculation. The average height of an indicator diagram is 
called its mean effective pressure, abbreviated M.E.P. Multiply- 
ing together the M.E.P. in pounds per square inch, the length of 
the stroke in feet, and the area of the piston in square inches will 
give the number of foot-pounds of work performed during the 
time in which the diagram was made, or one revolution. Multi- 
plying the above product by the number of revolutions per 
minute will give the number of foot-pounds of work performed 
per minute. This may be expressed in a formula as follows: 

W = Plan 
in which W = the number of foot-pounds of work per minute 
P = the M.E.P. in pounds per sq. in. 
I = the length of the stroke in feet 
a = the area of the piston in sq. in. 
n = the number of revolutions per minute (r.p.m.) 

Example. — A 20" X 24" engine makes 240 r.p.m. and the indicator 
diagrams show a M.E.P. of 63 lb. per sq. in. How many foot-pounds of 
work is the engine performing per minute? 

Solution. — The length of the stroke is 24" = 2 ft. 

The are of the piston is 

.7854 X 202 = 314.16 sq. in. 

Hence the work performed is 

W = Plan 

= 63 X 2 X 314.16 X 240 
= 9,500,198 ft.-lb. per min. 

The mean effective pressure may be measured directly from the 
indicator diagram by the method of ordinates or by means of an 
instrument called a planimeter. The method of ordinates con- 
sists in dividing the diagram into a number of parts and measur- 

98 



INDICATED AND BRAKE HORSEPOWER 



99 



ing its height at each of these parts, taking into account the 
scale to which the diagram is drawn, that is, the number of the 
indicator spring. 

The ordinate method of measuring the mean effective pressure 
is illustrated by Fig. 66. The limiting lines at right and left of 
the diagram are first drawn perpendicular to the atmospheric 
line, and the space between them divided into ten equal parts. 
Vertical lines, as shown at 1, 2, 3, 4, etc., running through the 
center of these spaces are then drawn and the length of each 




Fig. 66. 

between the upper and lower lines of the diagram is measured. 
Adding these lengths together, multiplying their sum by the 
scale of the indicator spring, and dividing by 10 will give the 
average pressure or M.E.P. 

To obtain the centers of the ten spaces previously mentioned, a 
convenient method is to take an ordinary scale and place it as 
shown in Fig. 66 so that the diagonal length between the limits 
of the diagram will be exactly 5 inches. Then at the left of the 
scale point off at 3^^ inch, and from there on every 3^^ inch towards 
the right of the diagram. The last point will be at 4^ inches. 
From these points draw vertical lines through the diagram per- 
pendicular to the atmospheric line. 



100 



STEAM ENGINES 



A convenient method of obtaining the combined lengths of the 
ordinates is to take a narrow strip of paper and mark on its edge 
the height of each ordinate. Begin with No. 1 and mark its 
length on the paper. Then place the mark made for No. 1 at 
the end of No. 2 and make a mark on the paper at the other end 
of No. 2 and so on until the combined length of the 10 ordinates 
is obtained. Measure this combined length with an ordinary 
inch scale, multiply by the scale of the spring, and divide by 10. 
The result will be the M.E.P. for the diagram. For example, 
suppose the combined length of the 10 ordinates in Fig. 66 meas- 
ures 7.8 inches and the diagram was drawn with 
a No. 60 spring. The M.E.P. would then be: 

7.8 X 60 . _ _ ,, 

— Yn — ^ ^"-^ ^^- P^^ ^^- ^^• 

In a noncondensing engine having insufficient 
load where the steam is cut off at a very early 
part of the stroke, a diagram may be obtained 

similar to the one shown 

■^JJJ__J__ ~ ~ "^J^ the diagram must be treated 
as two distinct parts, the 
loop at the toe of the dia- 
gram near the end of expansion being treated as negative and 
the other part being treated as positive. The whole diagram is, 
in this case, divided into 20 equal parts. From the combined 
lengths of the positive ordinates which number from 1 to 8, 
subtract the combined lengths of the negative ordinates, which 
number from 9 to 20, then multiply the difference by the 
scale of the spring and divide by 20, which is the number of 
ordinates. The result will be the mean effective pressure for the 
whole diagram. For example, if the combined length of the 
positive ordinates is 4 inches and the combined length of the 
negative ordinates 1.7 inches, the difference would be 2.3 inches. 
If the scale of the spring is 60, the M.E.P. would be 

2.3 X 60 

per sq. 




Fig. 67. 



20 



= 6.8 lb. 



m. 



which is the average pressure for the entire diagram. 

The ordinate method of measuring the mean effective pressure, 
described above, gives only approximate results. The approxi- 
mation is closer the larger the number of parts into which the 



INDICATED AND BRAKE HORSEPOWER 



101 



diagram is divided, but 10 divisions give results which are close 
enough for most practical purposes. The most accurate and 
quickest method of measuring the mean effective pressure is by 
means of a planimeter and this method should be used if a plani- 
meter is at hand. 

There are several makes of planimeters on the market, alike in 
principle but differing in details of construction. The one shown 
in Fig. 68 will be found very convenient for measuring mean 
effective pressure from indicator diagrams. This instrument 
consists of two arms AB and CD which are pivoted so they may 
move with respect to each other. In preparing the instrument 
for use the two points E and F on the back of the arm CD are 
set a distance apart equal to the length of the diagram. This 
adjustment should be made as close as possible by hand and a 




Fig. 68. 

final and closer adjustment made by means of the screw G. The 
two arms of the planimeter are then held at approximately 90° 
to each other, the point H is placed near the center of the indi- 
cator diagram, and the point J is pressed firmly into the board 
to hold it stationary, the small weight K being placed on it for 
the same purpose. The point H is next placed at one corner of 
the diagram, preferably the upper left-hand corner, and the small 
wheel M turned until its zero is opposite the zero on the fixed 
scale. The instrument is now in position for measuring the 
M.E.P. of the diagram. This is done by tracing out the diagram 
with the point H, following around the diagram in a clockwise 
direction. When the point H has been moved entirely around the 
diagram and brought back to its starting point, the number on 
the wheel opposite the zero point of the fixed scale is read. This 
number, when multiplied by the scale of the spring and divided 
by a constant which is usually 40, gives the M.E.P. of the diagram 
in pounds per sq. in. In case there is a loop in the toe of the 
diagram, as in Fig. 67 the point is first carried down the expansion 



102 STEAM ENGINES 

line and then around the loop in a counter-clockwise direction. 
This automatically subtracts the average pressure of the loop 
from the average pressure of the remainder of the diagram. 

Indicated Horsepower. — The rate at which work is performed 
in the engine cylinder, which is calculated from the indicator 
diagram, is called the indicated horsepower abbreviated I.H.P. 
After the M.E.P. of the indicator card has been measured, the 
indicated horsepower may be calculated by the formula: 

THP = ^-^ 
i.xi.r. 33Q00 

in which I.H.P. is the indicated horsepower 

P is the M.E.P. in pounds per sq. in. 

I is the length of stroke in feet 

a is the area of the piston in sq. in. 

n is the number of revolutions per minute 

The above formula gives the I.H.P. from a single indicator dia- 
gram, which is taken from but one end of the cylinder, hence, 
to find the total I.H.P. for the engine, the I.H.P. must be calcu- 
lated for each end of each cylinder and their sum taken. It should 
be remembered that the piston rod occupies a portion of the 
area of the piston, and, for accurate results, its area must be 
subtracted from the area of the piston when calculating the 
I.H.P. for the crank end of the cylinder. 

Example. — Calculate the indicated horsepower of a 20" X 24" simple 
engine, running at 240 r.p.m., the M.E.P. for the head end of the cylinder 
being 48 lbs. per sq. in. and for the crank end 49 lbs. per sq. in. The 
diameter of the piston rod is 2}^ inches. 

Solution. — The area of the piston on the head end is 

.7854 X 202 = 314.16 sq. in. 
The area of the piston rod is 

.7854 X 2.52 = 4.91 sq. in. 
The area of the piston on the crank end is 

314.16 - 4.91 = 309.25 sq. in. 
The length of stroke is 

24 

j-2 = 2ft. 

The indicated horsepower for the head end is 

Plan _ 48 X 2 X 314.16 X 240 
• 33,000 33,000 

The indicated horsepower for the crank end is 

P{an ^ 49 X 2 X 309.25 X 240 
l.M.r. 33Q00 33,000 



INDICATED AND BRAKE HORSEPOWER 103 

The total I.H.P. is 

219.3 + 220.4 = 439.7 

The indicated horsepower may be calculated approximately by using the 
average M.E.P. for the two ends of the cylinder and neglecting the area of 
the piston rod. The total indicated horsepower is then calculated by the 
formula: 

TTTP - 9 ^^^^ 

Applying this formula to the above example would give 
TTTP r^ Plan 48.5X2X314.16X240 

The result is 3.5 horsepower larger than the result obtained by the other 
method of calculation. 

Engine Constant. — In the formula for indicated horsepower 
it will be noted that, for any particular engine, the part of the 
formula 

la 
33,000 

is a constant quantity. This quantity is called the engine 
constant. Since the net area of the piston will be different on the 
head and crank ends, an engine will have a head end engine 
constant and a crank end engine constant. When the engine 
constant has once been calculated, the horsepower may be 
found at any time by observing the M.E.P. and speed and multi- 
plying these quantities by the engine constant. 

Brake Horsepower. — The indicated horsepower is the power 
developed in the cylinder of an engine. This power is trans- 
mitted through the piston rod, crosshead, connecting rod, crank 
and main shaft to the flywheel and a portion of it is lost by friction 
in the various bearings of the engine. Hence the amount of 
power delivered at the flywheel will be smaller than that developed 
in the cylinder. The power delivered to the flywheel is called 
the brake horsepower, abbreviated B.H.P. It may be measured 
by means of a device called a friction brake, hence the name brake 
horsepower. 

A friction brake usually consists of a band which is clamped 
on the face of the flywheel and which may be tightened so as to 
produce more or less friction between it and the flywheel. The 
power of the engine is expended in overcoming the friction of the 
brake, which is arranged in such way that the pull of the engine 
upon the brake may be measured. The brake horsepower is 

10 



104 



STEAM ENGINES 



calculated from the pull and speed of the engine and the dimen- 
sions of the brake. 

A common form of friction brake, called a Prony brake, is 
shown in Fig. 69. This brake consists of a wooden beam C and a 
band B made of a number of hard wood blocks fastened to a 
sheet iron band and passing around the flywheel A. The beam C 
has a steel knife-edge fastened to its under side near the end and 
resting on an iron plate on top of the stand E. The stand E 
rests on a platform scale so the pull of the engine upon the brake 
may be weighed. One end of the band containing the friction 
blocks is fastened to the beam by passing through it and having 




Fig. 69. 

a nut on its end. The other end of the band is held by hand 
wheel D so it may be tightened and its friction adjusted. The 
edges of the flywheel form inwardly projecting flanges so that a 
stream of water may be run into the flywheel to keep it cool. 
Preparatory to using the brake, the distance L from the center 
of the brake to the knife-edge is measured, the stand E is weighed, 
and the unbalanced weight of the brake about the center line 
FG is obtained. This may be done by suspending the brake by a 
cord at the point F while the end of the beam rests on a scale 
and noting its weight. In using the brake the engine is brought 
up to full speed and the band tightened as much as possible 
without reducing the speed. The weight registered on the scales 
and the speed of the engine are observed at the same time. 



INDICATED AND BRAKE HORSEPOWER 



105 



The brake horsepower may now be calculated by the formula: 

2TrlnW 



B.H.P. = 



33,000 



in which 



B.H.P. is the brake horsepower 
I is the length from center of flywheel to knife- 
edge in feet 
n is the number of revolutions per minute 
W is the pull of the engine on the brake, in pounds 
W = weight registered on scales minus weight 

of stand minus unbalanced weight of brake 
TT = 3.1416 




Fig. 70. 

Example. — What is the brake horsepower of a steam engine running at 
210 r.p.m. when fitted with a Prony brake w^hich measures 8 feet from 
center of the flywheel to the point of support at the end of the arm, the 
scale reading 742 lbs., the unbalanced weight of the brake being 13 lbs., and 
the weight of the standard being 10 lbs.? 



Solution 



Then 
B.H.P. = 



TT = 742 - 13 - 10 = 729 lb. 
n = 210 r.p.m. 
Z = 8 ft. 

2TvlnW 2X3.1416X8X210X719 



33,000 



33,000 



= 230 horsepower 



106 STEAM ENGINES 

Another form of brake for measuring power is shown in Fig. 
70. This form of brake is called a rope brake because the friction 
which furnishes the load for the engine is produced by a rope 
wound around the flywheel. In this brake the ends of the rope 
are attached to the top crosspiece C of a wooden frame which 
rests on a platform scales. The rope is looped around the 
flywheel and the middle attached to a screw which passes through 
the bottom crosspiece C. This screw passes through a hand 
wheel which is used to tighten the rope and thus regulate the 
load on the engine. Instead of a hand wheel a large nut may be 
used for this purpose. 

The brake horsepower, as measured with this form of brake, 
may be calculated from the formula: 

■R TT p ^ '^T^RWri 
33,000 

in which R = the radius of brake or distance from center of 

wheel to center of rope, in feet 

W = Pull of the engine = weight indicated by 
scales minus weight of wooden frame, in 
pounds 

n = number of revolutions per minute 

TT = 3.1416 

Mechanical Efficiency. — The mechanical efficiency of an engine 
or its efficiency considered simply as a machine, is the ratio of the 
brake horsepower to the indicated horsepower or 

"D TT p 

Mechanical Efficiency = ^ tt p 

This quantity is always less than one, since there is always a loss of 
power by friction in the engine; that is, the brake horsepower is 
always less than the indicated horsepower. The mechanical 
efficiency of steam engines varies from 85 per cent, to 95 per cent. 
The difference between the indicated horsepower and the 
brake horsepower is the amount of power required to overcome 
the friction. This quantity is sometimes called the friction 
horsepower or 

Friction horsepower = I.H.P. — B.H.P. 

The friction horsepower is the indicated horsepower of the 
engine when it is running without load. 



CHAPTER VIII 

ACTION OF STEAM IN THE CYLINDER 

Cylinder Condensation. — It is a well-known fact that the steam 
engine is a wasteful machine for developing power because it 
turns into work only a small part of the heat energy delivered to 
it. The amount of work obtained from a steam engine is often 
only 4 or 5 per cent, of the amount of energy delivered to it, and 
it rarely exceeds 20 per cent. This means that from 80 to 96 
per cent, of the heat energy supplied to the steam engine is 
wasted or at least is not utilized. For example, suppose an 
engine uses 35 pounds of dry saturated steam per hour for each 
indicated horsepower developed, the steam having an absolute 
pressure of 100 lb. per sq. in. The heat delivered to the engine 
amounts to 35 X 1186.3 = 41,520.5 B.t.u. per horsepower 
per hour and from this amount of energy only one horsepower is 
obtained. One horsepower for an hour is equivalent to 2545 
B.t.u., therefore, the part of the energy supplied which is 
turned into work is only 

... -or> r = .0613 or'6.13 per cent., 

the remaining 93.87 per cent, being lost or wasted. A large part 
of this loss occurs through the exhaust but another considerable 
part occurs through the condensation of steam in the cylinder. 

Cylinder condensation is caused by the alternate cooling and 
heating of the cylinder walls as they are alternately in contact 
with high pressure steam (which has a high temperature) during 
admission, and to low pressure steam (which has a lower tempera- 
ture) during exhaust. 

The exchanges of heat taking place between the steam and 
cylinder walls may best be studied by considering the cycle of 
events occurring in only one end of the cylinder. Exhaust occurs 
during the greater part of the return stroke of the piston and 
during this time the cylinder walls, face of the piston, and 
cylinder head are in contact with steam having a comparatively 
u 107 



108 STEAM ENGINES 

low temperature, thus cooling these parts of the engine. Com- 
pression at the end of the return stroke raises somewhat the 
temperature of the steam in the clearance space, but the warming 
effect on the cyUnder is small because the temperature of the 
compressed steam is not so high as the steam admitted from the 
boiler and the piston is near the end of its stroke, exposing very 
little of the cylinder walls to the compressed steam. Most of 
the surface so exposed consists of the face of the piston and the 
cylinder head. Consequently, when the admission valve opens 
and a fresh charge of high pressure (and high temperature) 
steam rushes into the cj^Hnder it meets comparatively cool metal 
surfaces and a part of it is condensed, collecting in a thin layer 
of water on these surfaces. As the piston advances on its 
forward stroke it uncovers more and more of the chilled cylinder 
walls which condense still more of the steam which is being 
admitted to the cylinder, with the result that, up to the point of 
cutoff, from 30 to 50 per cent, of the steam fed into the cylinder 
during admission is condensed, thus requiring that a greater 
volume of steam be supphed to the cylinder than if none of it was 
condensed. The condensation occurring up to the point of 
cut-off is called initial condensation. 

After cut-off the piston continues to advance and uncover more 
of the cooled cyhnder walls. Hence, condensation continues 
after cut-off but at a lessening rate because, after cut-off, the 
steam in the cylinder is expanding and its temperature falling. 
The difference in temperature between the steam and the 
cylinder walls is not so great. Besides the condensation due to 
contact between the steam and cooler cylinder walls there is 
now also a certain amount of condensation caused by energy being 
taken out of the expanding steam to move the piston. 

Whatever steam is condensed during the early part of the 
stroke is deposited in the form of a thin film of water on the 
cylinder walls, the face of the piston, and the cylinder head. This 
film of water has a temperature equal to that of the steam from 
which it was formed, hence it is at or very near the boiUng point 
corresponding to the pressure of the steam. After cut-off the 
steam in the cylinder begins to expand and its pressure falls. 
This lowers the boiling point below the temperature of the water 
already deposited on the inside of the cylinder with the result 
that this water begins to reevaporate. As expansion proceeds, 
the boiling point is lowered and the difference between the boihng 



ACTION OF STEAM IN THE CYLINDER 109 

point and the temperature of the layer of water becomes larger, 
therefore reevaporation proceeds at a faster and faster rate. 
For this reason, a point is reached soon after cut-off where the 
reevaporation balances the condensation and at this point the 
amount of water in the cylinder is a maximum. From this 
point on reevaporation occurs faster than condensation and the 
amount of water in the cylinder grows smaller. When release 
occurs, there is a sudden drop in pressure, accompanied by a 
sudden drop in the boiling point, and the layer of water on the 
cylinder walls reevaporates very fast. During exhaust the 
pressure remains low and reevaporation continues at a rapid 
rate, if there is still any water remaining in the cylinder. If the 
initial condensation has not been very great, however, the water 
may be all reevaporated at the beginning of exhaust, and the 
exhaust steam will then be dry. 

It might be thought that if the water in the cylinder is all re- 
evaporated no harm would be done It should be remembered, 
however, that any water reevaporated near the end of expansion 
is at a lower pressure than when condensed and consequently 
it cannot be expanded as much as if it had not 'been condensed 
but had remained in the form of steam and expanded through 
the whole range of pressure. 

If the steam admitted to the cylinder already contains some 
water, as would be the case if wet steam were supplied, the amount 
of water reevaporated during expansion and exhaust may be 
greater than the condensation during admission. This would 
cause a much larger quantity of heat to be taken from the cylinder 
walls and the chilling effects of reevaporation to be greatly in- 
creased. It is an advantage therefore to supply only perfectly 
dry steam to an engine in order to reduce the amount of water 
in the cylinder. 

In the above discussion of cylinder condensation, the events 
occurring in ony one end of the cylinder have been considered. 
If we consider these events as occurring in the head end, for 
instance, then the events occurring in the crank end have some 
influence upon the condensation and reevaporation in the head 
end. During the first part of exhaust from the head end, this 
end of the cylinder is in contact with low temperature steam 
and may be further cooled by reevaporation, but at the same 
time admission and expansion are occurring in the crank end and 
this end of the cylinder is being warmed shghtly by contact with 



110 STEAM ENGINES 

high temperature steam. Admission and expansion in the 
crank end, therefore, reduces sHghtly the coohng of the head end 
and hence reduces the amount of condensation that would occur 
in the head end. The reduction in initial condensation from 
this cause will depend upon the lateness of the cut-off in the 
opposite end of the cylinder, the condensation being less for a 
late cut-off than for an early one because a late cut-off exposes 
more of the cylinder walls to high temperature steam. 

The cooling effects of reevaporation depend upon the reduction 
in pressure of the steam during expansion, being greater the more 
fully the steam is expanded. The greatest amount of expansion 
occurs with an early cut-off, hence, an early cut-off increases 
reevaporation. It will thus be seen that the loss of steam by con- 
densation and the cooling of the cylinder by reevaporation are 
both increased by an early cut-off. 

The give and take of heat between the steam and cylinder walls 
does not affect all of the metal in the cylinder because the heat 
transfers occur so rapidly that their effects do not have time to 
extend very far into the metal. The outside of the cylinder 
assumes a temperature between that of the exhaust and the 
admission steam and this temperature remains practically 
constant while the engine is running. The inner surfaces of the 
cylinder and piston, however, experience great changes in tem- 
perature, being alternately heated and cooled, but such changes 
of temperature grow less the greater the distance from the inner 
surface at which they are measured. It is probable that the 
depth of metal thus affected does not average more than .02 
to .03 inch, and this depth is less for high rates of revolution 
than for low rates because there is not time, with a high rate of 
revolution, for the transfer of heat to take place. Other things 
being equal, the losses from cylinder condensation and re- 
evaporation are less for high speed engines than for low speed 
ones. 

It has been pointed out in a previous paragraph that the 
amount of heat taken from the cylinder walls by reevaporation 
exceeds the amount given to the walls by condensation. In some 
cases the amount of heat taken from the cylinder walls may be 
only slightly greater than that given to them and, in such cases, 
it might be thought that the cooling effect would be very small. 
In this connection, however, it should be remembered that the 
transfer of heat affects only a thin layer of metal and the weight 



ACTION OF STEAM IN THE CYLINDER 111 

of metal affected, therefore, is small. Since one B.t.u. will 
change the temperature of 73^^ lbs. of cast iron one degree or one 
pound of cast iron 73-^ degrees it will be seen that the transfer 
of even a small amount of heat may produce comparatively 
great changes of temperature in the inner surfaces of the cylinder. 

The effects of high speed and late cut-off in reducing the losses 
from cylinder condensation and reevaporation have been noted. 
Other remedies that are used for accomplishing this purpose are : 
the use of superheated steam; the use of a steam jacket surround- 
ing the cylinder; and compounding, or dividing the total range 
of expansion between two or more cylinders. The effects of 
compounding will be considered in a later chapter. 

The benefits to be derived from the use of superheated steam 
come from the prevention of initial condensation. Superheated 
steam contains more heat per pound than saturated steam at the 
same pressure, and, before it can be condensed, the extra heat 
which it contains must first be taken out of it, thus reducing its 
temperature and changing it into saturated steam. Taking more 
heat from it will then cause condensation. When superheated 
steam is supplied to an engine the heat needed to warm the 
cooled cylinder walls may be supplied from the extra store of 
heat which the steam contains and there will be no initial con- 
densation. In order to fully accomplish this purpose, however, 
the steam must be superheated enough to secure dry steam at the 
point of cut-off. If it is not superheated to this degree all the 
excess heat will be taken from the steam before cut-off and it 
will then begin to condense and deposit moisture on the cylinder 
walls. 

There is but little advantage in using steam which is super- 
heated to such an extent that it will still be superheated after 
expansion commences, because reevaporation begins at this 
point and also because the condensation which occurs before cut- 
off causes more serious loss than that which occurs after cut-off, 
and it is, therefore, of more advantage to prevent the initial 
condensation. 

Since initial condensation is usually greatest in the plain slide 
valve type of engine, which employs a late cut-off, it is to be 
expected that more benefit may be derived from the use of super- 
heated steam in this type of engine than from those of other 
types, employing an earlier cut-off, or those making use of com- 
pound expansion. 



112 STEAM ENGINES 

A steam jacket consists of a hollow space surrounding the 
cylinder, connected to the main steam supply for the engine, and 
filled with steam at high pressure. The steam jacket is supposed 
to benefit by keeping the cylinder walls at a uniformly high 
temperature and preventing the rapid changes of temperature 
in the walls. It is found in practice, however, that but little 
benefit is obtained from the steam jacket because the changes of 
temperature take place in only a thin layer of metal on the inner 
surfaces of the cylinder and these changes of temperature occur 
so rapidly that the heat from the jacket does not have time to 
flow through the walls rapidly enough to prevent them. Cylinder 
condensation is reduced somewhat, however, by the presence of 
the jacket because it maintains a higher average temperature of 
the cylinder walls. On the other hand, it must be remembered 
that whatever heat is supplied by this means comes from the 
condensation of steam in the jacket and also that the presence of 
the jacket makes the outside diameter of the cylinder greater and 
increases its surface. The greater surface of the cylinder together 
with its higher temperature increases the amount of heat lost by 
radiation from the cylinder. Since the advantages of a steam 
jacket are doubtful and its presence increases the cost of the 
engine, it is not used as much now as formerly. Instead, the 
cylinders of the better classes of engines are now simply lagged 
with a nonconducting substance to reduce the radiation of heat. 

The Uniflow Engine. — Within recent years a single expansion 
engine has been designed with a view to reducing the losses from 
cylinder condensation. This type of engine is called the Uni- 
flow Engine. A section of the cylinder of the uniflow engine is 
shown in Fig. 71. The cylinder contains no exhaust valves but 
a ring of exhaust ports are cut in the middle of the cylinder 
which is uncovered by the piston at the end of its stroke so that, 
in effect, the piston is the exhaust valve. For this purpose, both 
the cylinder and the piston are made longer than in the ordinary 
steam engine. The two admission valves. A, which are of the 
Corliss type, are located in the cylinder heads and the steam 
spaces over the valves become steam jackets for the heads. The 
clearance pocket B is also kept filled with steam so that the head 
is completely steam jacketed. 

After cut-off, the steam expands behind the piston as in the 
ordinary types of engines, but at the end of expansion the piston 
uncovers for an instant the exhaust ports and the remaining 



ACTION OF STEAM IN THE CYLINDER 



113 



pressure in the cylinder falls. Oh the return stroke the steam 
at exhaust temperature and pressure is caught between the piston 
and cylinder head, and, as the piston moves back, this steam is 
compressed so that its temperature is gradually increased and 
at the end of the stroke the clearance space is filled with steam 
at admission pressure. The temperature of the steam in the 
clearance space is increased not only by compression but also by 
absorbing heat from the head jackets. The result is that the 
temperature of the steam in the clearance space may be raised 
even higher than that of the admission, therefore, when the ad- 
mission valve is opened the incoming steam meets no cold sur- 
faces and initial condensation is reduced. An indicator diagram 
from a uniflow engine is shown in Fig. 72. 




Fig. 71. 

By referring to Fig. 71 it will be seen that each end of the 
cylinder is provided with a large relief valve D opening into a 
pocket B in the cylinder head. This valve serves two purposes: 
First, it is a relief valve of large size which will relieve the engine 
of any entrained water ; second, if, when exhausting into a vacuum, 
the vacuum should be broken, it is necessary to provide the 
engine with a larger clearance volume in order to prevent exces- 
sive compression. These valves open automatically in case the 
vacuum is broken, and, if it is then desired to run the engine 
noncondensing, means are provided to back these valves off their 
seats, thus increasing the clearance space by the volume of the 
clearance pockets. The enlargements C, C, in the steam passages 
are to provide for expansion of the metal without distorting the 
cylinder. 



114 



STEAM ENGINES 



That these engines at least partly accomplish their purpose is 
shown by the fact that they have developed a horsepower with a 
steam consumption of only 13.2 pounds per hour, which is a 
very good performance for a single expansion engine even when 
exhausting into a vacuum of 26 inches. 

Measuring Cylinder Condensation. — The quantity of water 
present in the cylinder at any time between cut-off and release 
may be found from the indicator diagram and a knowledge of the 
weight of steam used by the engine. The method of doing this 
is best shown by means of an actual example. Fig. 73 shows an 
indicator diagram from the head end of a 12'' X 24'' engine, 
with a clearance on the head end of 7.9 per cent. The engine 
was running 100 r.p.m. when the diagram was taken, and, by 
condensing and weighing the exhaust for an hour, it was found 



/So/i-£fp f^f^es>s>u^£. 




Fig. 72. 



that the engine was using 2384.4 pounds of steam per hour. The 
barometer read 28.52 inches, equivalent to an atmospheric 
pressure of 14 lbs. per sq. in. The spring in the indicator was 
No. 100. By drawing on the diagram the ''dry steam line" 
SS, as expalined below, the percentage of water in the cylinder, 
or ''quality" of the steam at anytime between cut-off and release, 
may be measured directly from the diagram. Thus in Fig. 73, 
the. quality of the steam at cut-off is found by taking the pro- 
portion between the lengths of the lines AB and AC. From the 
diagram, the length of the line AB is 1.0 inch and the length of 
AC is 1.43 inches, therefore the quality of the steam at cut-off is 

AB 



AC "^ 1.43 



= .70 = 70 per cent. 



or, of the mixture in the cylinder at cut-off, 70 per cent, is dry 
steam and 30 per cent, is water. In other words, the initial con- 



ACTION OF STEAM IN THE CYLINDER 



115 



densation has been 30 per cent, of the steam suppHed to the 
cyUnder. In a similar manner the quahty of the steam at any 
other point F in the expansion hne is found by measuring the 
line DE, which is 2.28 inches and the line DF, which is 2.86 
inches and taking the ratio 



DE 
DF 



2.2^ 
2.86 



= .797 or 79.9 per cent. 



showing that the steam in the cylinder is dryer at the point E 
than it is at the point B, a result that is to be expected. 

In order to draw the dry steam line it is first necessary to 
locate the line of no pressure, OG, and the line of no volume, OH. 



H 




S 
\P 


^ 


E ^"^ 




A 

D 

L 







^---^F 


J 






III i 



M 



Fig. 73. 



The line OG is used as a base line from which to measure absolute 
pressures and is drawn parallel to the atmospheric line, J J, and at 
a distance below it equal to the atmospheric pressure, 14 lb. 
per sq. in. to the same scale to which the diagram is drawn. 
Volume of steam in the cylinder is also measured on this line. 
The line, OH, is used for determining the amount of steam in the 
clearance volume during compression and also to measure pres- 
sures. The line OH is drawn perpendicular to the atmospheric 
line and at a distance from the end of the diagram equal to the 
volume of the clearance space, to the same scale to which the 
diagram is drawn. By drawing limiting lines at the ends of the 
diagram, extending to the atmospheric line, the length of the 
diagram is measured on the atmospheric line and is found to be 



116 STEAM ENGINES 

3.34 in. The piston displacement is .7854 X P X 2 = 1.5708 
cu. ft. Therefore one inch length on the diagram represents a 

1.5708 
volume of ^"oj- = -47 cu. ft. The volume of the clearance 

is .079 X 1.5708 = .124 cu. ft. Therefore the line OH is laid 

.124 

off from the end of the diagram a distance of '~t^ = .264 inch. 

* .47 

The weight of steam in the clearance space during compression 
is found by taking any point such as K on the compression curve 
after the exhaust valve is closed and measuring the pressure 
and volume represented by this point. The absolute pressure of 
the point K is found by measurement to be 40 lb. per sq. in. 
The volume of one pound of dry steam at this pressure is, from the 
steam table, 10.3 cu. ft. The volume of steam in the clearance 
space is represented by the line LK which is .34 inch. It, there- 
fore, represents a volume of .34 X .47 = .1598 cu. ft., and its 

1598 
weight is ' ^ = .0152 pound. The weight of steam fed to each 

^ f^u V A u • 2384.4 ,,^ooiu 1192.2 

end ot the cyhnder per hour is — ^ — = 1192.2 lbs. or — w^ = 

19.87 lbs. per minute. Since the engine was running 100 r.p.m., 
the weight of steam fed to the engine while the diagram was being 

19 87 
drawn was -^rw^r = .1987 pound. The weight of steam expand- 
ing in the cyhnder each time was, therefore, .0152 + .1987 = 
.2139 pound and it is for this weight of steam that the dry steam 
line SS must be drawn. 

The dry steam line SS is drawn by taking from the steam table 
the volumes of one pound of dry steam at various pressures and 
multiplying them by the weight of steam expanding in the 
cylinder .2139 lb. Thus at an absolute pressure of 150 lb. per 
sq. in. the volume of one pound of dry steam is, from the steam 
table, 2.978 cu. ft., therefore, the volume of .2139 lbs. is .2139 X 
2.978 = .6443 cu. ft. Measuring from a distance equal to 

.6443 
.J = 1.37 inches the point Q is located, which represents 

a volume of .6443 cu. ft. Drawing the line PQ perpendicular 
to OG and of a length equal to 150 lbs. per sq. in., the point 
P is located, which is one point on the dry steam line. At an 
absolute pressure of 120 lb. per sq. in., the volume of one pound 
of dry steam is, from the steam table, 3.726 cu. ft., therefore, the 
volume of .2139 pounds is .2139 X 3.726 = .7970 cu. ft. Meas- 



ACTION OF STEAM IN THE CYLINDER 117 

.7970 
uring from a distance equal to ^"Ty~ = 1-70 inches, the point 

S is located, which represents a volume of .7970 cu. ft. Drawing 
the line RS perpendicular to OG and of length equal to 120 lb. 
per sq. in. the point R is located which is another point on the 
dry steam line. In a similar manner any number of points on 
the dry steam line may be found, and a smooth curve drawm 
through these will give the dry steam line SS. The quality 
of steam in the cylinder may then be measured from this line, 
as explained before. In order for this method of finding the 
quality of steam in the cylinder to give accurate results, there 
must be no leakage in the cylinder, as this would change the shape 
of the expansion line. 



CHAPTER IX 
STEAM ENGINE TESTING 

Principles. — The usual steam engine test is made to determine 
the weight of steam or the number of heat units which the engine 
consumes per hour for each horsepower developed or else to 
determine the efficiency of the engine. If possible, the steam 
consumption and efficiency should be based on the brake horse- 
power of the engine because this is the useful power of the engine. 
However it is sometimes impossible or impractical to obtain the 
brake horsepower and, in this case the steam consumption, and 
efficiency are based on the indicated horsepower. 

As practically all engines operate under variable loads, it is 
advisable in testing them, to test at different per cents of the full 
load in order to determine what performance may be expected 
under different conditions. For this purpose it is convenient 
to test an engine under one quarter load, one half load, three 
quarters load, full load, and one and one quarter of its full load 
capacity. With this data at hand a curve may be plotted with 
brake horsepower or indicated horsepower on one axis and steam 
consumption or efficiency on the other axis, and from the curve 
so obtained, one may determine what performance to expect 
from the engine under the load at which it operates most of the 
time. 

If the engine is belted and not very large, its brake horse- 
power may be measured with any of the forms of friction 
brakes described in a previous chapter, although these require 
the use of a special pulley designed to hold water for cooling. 
If the engine is connected directly to an electric generator, it 
will be necessary to determine separately the amount of power 
required to run the generator without load, or the friction load 
of the generator, so that this may be deducted from the out- 
put of the generator in calculating the brake horsepower of 
the engine. If the brake horsepower cannot be determined by 

118 



STEAM ENGINE TESTING 119 

one of these methods, it will be necessary to base the calcula- 
tions for steam consumption and efficiency upon the indicated 
horsepower. 

Steam Consumption. — The steam consumption is best deter- 
mined by means of a surface condenser. In this case the exhaust 
steam from the engine is simply run into a surface condenser 
where it is condensed and the condensate weighed. In order to 
secure accurate results by this method the condenser should be 
free from leaks and the condensate should be cooled to a tempera- 
ture that will prevent its giving off much vapor, as otherwise 
the loss of condensate by evaporation will seriously affect the 
results of the test. 

The steam consumption may also be determined by means of 
one of the commercial forms of steam meters which measures 
the weight of steam passing through it. If this method of 
determining the steam consumption is used, the steam meter 
should first be carefully calibrated to insure its giving accurate 
results. 

If neither of the above methods is available, it may be possible 
to isolate the boiler or boilers supplying the engine, so that all of 
the steam generated by the boiler is used in the engine. The 
feed water supplied to the boiler may then be measured and taken 
as the steam consumption of the engine. Sometimes it is 
necessary for the boiler used in this way to supply steam for some 
auxiliaries such as feed pumps, etc. In such a case it is necessary 
to determine separately the amount of steam used by the 
auxiliaries. This may usually be done, by condensing the 
exhaust steam from them. In using this method of determining 
the steam consumption of an engine extreme care should be 
taken to insure that there are no leaks, especially at branches 
stopped by valves. This is done by closing all valves in branches 
and the main stop valve at the engine so that the main 
supply pipe is open from the boiler to the engine valve, 
but closed everywhere else. With a quiet furnace fire so 
that there is no active evaporation the level of the water in 
the boiler is noted from time to time. If the water level falls, 
leakage is taking place and the leaks should be located and 
stopped or else the rate of leakage allowed for in the steam 
consumption. 

Steam Consumption from Diagram. — It is sometimes impos- 
sible to find the weight of steam used by an engine by condensing 



120 STEAM ENGINES 

the exhaust and weighing it, or by isolating the boiler and 
weighing the feed water. In such cases the weight of steam used 
may be found from the indicator diagram, but this method should 
not be used except when the weight of steam used cannot be 
found by any other method, because it is subject to serious 
errors on account of leakage of steam into or out of the cylinder 
and from one side of the piston to the other. 

The method of finding the steam consumption from the indi- 
cator diagram may be illustrated by Fig. 73. By this method 
it is first necessary to draw the no volume line, OH, and the no 
pressure line, OG, as described before. If the barometer reading 
is not known, it is customary to draw the line OG at a distance 
below the atmospheric line equal to 14.7 lb. per sq. in. to the 
same scale as the spring used in drawing the diagram. A line 
DE is drawn across the diagram parallel to the atmospheric 
line and at a point near the end of the expansion line. The 
weight of steam represented by the volume DE is the weight which 
is expanding in the cylinder. This weight minus the weight 
of steam compressed into the clearance space is the weight of 
steam fed to the cylinder at each stroke. The length of the line 
DE is 2.66 inches and since a length of one inch on the diagram 
represents a volume of .47 cu. ft. the volume represented by the 
line DE is .47 X 2.26 = 1.0622 cu. ft. At the point E on the 
expansion line, the steam in the cylinder has an absolute pressure 
of 65 lb. per sq. in. From the steam table one cubic foot of 
steam at 65 lbs. absolute pressure weighs .1503 pound, hence the 
weight of steam expanding in the cylinder is 1.0622 X .1503 = 
.1626 lb. 

The weight of steam in the clearance space is found by selecting 
a point K on the compression curve after the exhaust valve has 
closed and measuring the pressure and volume of steam repre- 
sented by this point. This was done before in drawing the dry 
steam line, and it was found that the weight of steam in the 
clearance space was .0152 pound. Therefore the weight of 
steam fed to the cylinder at each stroke was .1626 — .0152 
= .1474 lbs. Since the engine was making 200 strokes per min- 
ute, the weight of steam used per hour as shown by the diagram 
was .1474 X 200 X 60 = 1768.8 lbs. This weight makes no 
allowance for the condensation in the cylinder, hence it must 
be corrected by means of the values given in the following 
table. 



STEAM ENGINE TESTING 



121 



Percentage of 


Part of steam accounted for by the indicator diagram 


strokes completed 
at cut-off 


Simple engines 


Compound engines 
H.P. cylinder 


Triple expansion 

engines H.P. 

cylinder 


5 

10 
15 
20 
30 
40 
50 


0.58 . 

0.66 

0.71 

0.74 

0.78 

0.82 

0.86 


0.74 
0.76 

0.78 
0.82 
0.85 
0.88 


0.78 
0.80 
0.84 
0.87 
0.90 



By measurement it is found that cut-off in the above example 
occurs at about 19 per cent, of the stroke, hence the weight of 
steam from the diagram should be divided by .74 or 

1768 8 

— ;—— = 2390 pounds per hour as the probable weight of steam 

used by the engine. While this result is close to the actual weight 
of steam used, 2384 lbs., it must be remembered that this method 
of finding the weight of steam used is liable to serious error, 
sometimes amounting to as much as 50 per cent. 

It is customary to express the steam consumption per horse- 
power hours in terms of the ''dry steam equivalent," that is, in 
terms of the number of pounds of dry steam which would contain 
as many B.t.u. as is contained by the steam of the quality 
actually supplied to the engine. In order to make this calcula- 
tion, the quality of steam supplied to the engine during the test 
is measured and the number of heat units in one pound of this 
steam calculated. This quantity multiplied by the number 
of pounds of steam supplied per horsepower gives the total 
number of heat units supplied to the engine per horsepower. 
The number of heat units per horsepower is then divided by the 
number of heat units in one pound of dry steam of the same 
pressure as that supplied to the engine and the quotient will be 
the number of pounds of equivalent dry steam supplied per 
horsepower. 

This method of stating the performance of an engine forms a 
very satisfactory basis for comparing one engine with another, 
provided the engines are operating under similar conditions, 
but the quality of the engine cannot be judged by this method 
of comparison if one engine uses superheated steam and the 



122 



STEAM ENGINES 



other one uses saturated steam. Neither does a comparison of 
efficiencies as calculated in a previous paragraph form a satis- 
factory basis for comparing the qualities of the engines if one is 
run condensing and the other noncondensing. The steam con- 
sumption of engines varies widely, depending upon the kind of 
engine and the conditions under which it is operated. Some 
of the best performances of engines that have been recorded are 
given below. 



c!„^«^ Lbs. of 

hell' dry steam 
degrees Pp^j/iS^r' 



H.P. 



Gage 

pressure 

lbs. 



Vacuum 
inches 



R.P.M. 



Westinghouse vertical at 

Brooklyn, N. Y 5,400 



185 



27.3 



76 



Rockwood-Wheelock at Natick, 

R.I 595 



Mcintosh & Seymour at 

Webster, Mass 1,076 



Rice & Sargent at Brooklyn, 
N. Y 



627 



Rice & Sargent at Philadelphia, 
Pa 



420 



159 



25.4 



123 



27.10 



151 



28.6 



142 



25.8 



Horizontal 4-valve. 



Leavitt pumping engine at 
Chestnut Hill, Mass 



658 



575.7 



150.4 



26.4 



175.7 



27.25 



76.4 



99.6 



11.93 



20 



13.0 



121 



102 



80 



297 



16.4 



12.76 



12.10 



9.56 



12.03 



50.6 



11.20 



Duration of Engine Test. — The duration of a test will depend 
upon the conditions under which the test is conducted and upon 
the methods used in making the different measurements. If 
the engine is tested under a brake load which may be kept con- 
stant, the test is simplified and the time of conducting the test 
shortened. During each test at a constant load, the steam used 
is weighed at uniform intervals of time, say ten or fifteen min- 
utes. When there are six or eight of these which are nearly 
constant in amount, the run may be discontinued provided the 
error of starting and stopping is not large. The error of starting 
and stopping will depend upon the method used in measuring 
the steam consumption. If the steam consumption is measured 
from a surface condenser, the error from starting and stopping 
will be only the difference in the amounts of condensate in the 
condenser at the beginning and end of the test. This will be a 



STEAM ENGINE TESTING 123 

relatively small amount. With a steam meter used for measuring 
the steam consumption the error of starting and stopping the 
test will be even smaller than with a surface condenser. But, 
when the steam consumption is determined by measuring the 
feed water supplied to a boiler, the error of starting and stopping 
the test will be large and the test must be conducted for a greater 
length of time. The error in this case may arise from a difference 
in level of the water in the boiler at starting and stopping or from 
a difference in the density of the water due to a different rate of 
boihng. 

The frequency of taking indicator diagrams from the engine 
will depend on how the load is varying. With a constant brake 
load indicator diagrams may be taken at ten minute intervals, 
but with a varying load the intervals should be from three to 
five minutes. The object in any case is to get a fair average of 
the mean effective pressure or indicated horsepower. 

Efficiency of Steam Engines. — The term efficiency usually 
means the ratio between the work obtained from a machine and 
the energy supplied to it. The efficiency of a steam engine may 
therefore be expressed as: 

-pffi • _ Work obtained from the engine 

•^ Energy supplied to the engine 

On this basis the efficiency of a steam engine is very low on ac- 
count of the large losses of heat taking place in the engine itself 
and of the large amount of heat rejected by the engine in the 
exhaust steam. 

In calculating the efficiency of a steam engine either the brake 
horsepower or the indicated horsepower may be used as the ^'work 
obtained from the engine," but it should be stated which of these 
is used. If the energy supplied to the engine is expressed in 
B.t.u. per minute, the work obtained from the engine should 
also be expressed in B.t.u. per minute; this may be done by 
multiplying the horsepower by 42.42. If the energy supplied 
to the engine is expressed in B.t.u. per hour, the horsepower 
should be multiplied by 2545 to obtain the work done by the 
engine per hour in B.t.u. 

The ''energy supplied to the engine'' should include all of the 
heat actually supplied to the engine, calculated above the heat of 
the liquid for the exhaust pressure. It would not be fair to the 
engine to charge it with the heat of the liquid below the exhaust 

12 



124 STEAM ENGINES 

pressure because the engine could not possibly change this heat 
into work. Neither should the engine be charged with all of 
the heat in the steam above 32°F., because the engine would 
have to exhaust into a vacuum in which the absolute pressure 
was only .089 lb. per sq. in. in order to make all of this heat 
available, and it is not possible to produce and maintain this 
low pressure in a condenser. 

The above expression for efficiency of a steam engine now be- 
comes : 

B.H.P. X 42.42 
^ ~ W{qL + h-hi) 
in which 

E is efficiency of the engine 

B.H.P. is the brake horsepower of the engine. (Use 

I.H.P. if necessary.) 
W is the weight of steam supplied to the engine per minute 
q is the quality of the steam supplied to the engine 
L is the latent heat of the steam per pound at admission 

pressure 
h is the heat of the liquid per pound at admission pressure 
hi is the heat of the liquid per pound at exhaust pressure 

Example. — An engine receiving steam at a pressure of 150 lbs. per sq. in. 
absolute and having a quality of 98 per cent, develops 600 I.H.P. and uses 
12,000 lbs. of steam per hour. The exhaust pressure is 16 lb. per sq. in. 
absolute. What is the efficiency of the engine? 

Solution. — The weight of steam used per minute is 

12,000 _.-„ 
— x^r— = 200 lbs. 
60 

The latent heat of steam at 150 lbs. per sq. in. absolute pressure is 863.2 
B.t.u. 

The heat of the liquid at 150 lbs. per sq. in. absolute pressure is 320.2 
B.t.u. 

The heat of the liquid at 16 lb. per sq. in. absolute pressure is 184.4 B.t.u. 

Therefore 

I.H.P. X 42.42 
^ "~ W{qL + h -hi) 

_ 600 X 42.42 

~ 200(.98 X 863.2 + 330.2 - 184.4) 
^ 26,452 _ 25,452 
" 200 X 991.8 ~ 198,346 
= .1313 or 13.13 per cent. 

The above formula for calculating the efficiency of a steam 
engine is used only when the engine is supplied with saturated 



STEAM ENGINE TESTING 125 

steam. If superheated steam is supplied, the heat supphed to 
the engine should include the heat required to superheat the 
steam. 

Efficiency of a Perfect Engine. — The efficiency of an imaginary 
perfect engine may be calculated by the formula 

in which Ep = the efficiency of the perfect engine 

Ti = the absolute temperature of the admission steam 
T2 = the absolute temperature of the exhaust steam 

By absolute temperature is meant the temperature reckoned from 
the absolute zero of temperature or the point below which it 
would be impossible to cool any substance. This point is located 
at 460° below zero on the Fahrenheit scale of temperatures, 
hence to change Fahrenheit temperature to absolute temperature 
it is necessary to add 460° to the Fahrenheit temperature. 
If the engine in the above example had been a perfect engine, 
its efficiency would have been calculated as follows: The tem- 
perature of steam at 150 lb. per sq. in. absolute pressure is, from 
the steam table, 358. 5°F. and its absolute temperature is, there- 
fore, 358.2 + 460 = 818.5°. The temperature of steam at 16 
lb. per sq. in. absolute pressure is, from the steam table, 216.3° 
F and its absolute pressure is, therefore, 216.3° + 460 = 676.3°. 
Hence, the efficiency of the perfect engine would be 

^ 818.5-676.3 ,_, ,^^, 

^p = ^T^~K = -1734 or 17.34 per cent. 

oio.O 

It will be observed that the efficiency of the perfect engine 
depends only upon the temperature of the admission steam and 
of the exhaust steam. It follows, therefore, that the efficiency 
of the perfect engine can never be 100 per cent, because the 
maximum temperature of the admission steam is limited, and 
also it is impossible to reduce the temperature of the exhaust 
steam to the absolute zero. 

The efficiency of a perfect engine, as calculated above, is often 
used as a standard by which to compare the efficiencies of dif- 
ferent engines, the admission and exhaust temperature being 
taken the same for both the perfect and the actual engine. The 
method of comparison is to divide the efficiency of the actual 
engine by the efficiency of the perfect engine, the result being 
called the efficiency ratio or 



126 STEAM ENGINES 

E 
Efficiency Ratio = ^^ 

Jiip 

in which E = the efficiency of the actual engine 

Ep = the efficiency of the perfect engine taken between 
the same hmits of temperature 
The efficiency ratio for the engine mentioned in the example 

above is 

E 1 '^ 1 *^ 
Efficiency Ratio = ^^ = ^^-^ . = .7514 or 75.14 per cent. 

Jbp l/.o4 

Computations. — In order to show how the computations for 
an engine test are made and also to show a form for reporting 
the results, the data and computed results of an efficiency test 
of an engine are given below. Both the data taken in making 
the test and also the results computed from this data are first 
tabulated, the calculated results being in heavy face type, and 
following this the method of making the calculations is shown. 

The test given below is one of a series of tests made on an auto- 
matic high speed engine in the steam laboratory of the University 
of Wisconsin to determine its economy, thermal efficiency, 
and mechanical efficiency. During each test the engine carried a 
practically constant load made by a Prony brake similar to that 
shown in Fig. 69. The steam consumption was measured by 
passing the exhaust steam from the engine into a surface conden- 
ser operated at atmospheric pressure, where it was condensed 
and weighed. This method of loading the engine and deter- 
mining the steam consumption made it possible to secure suffi- 
ciently accurate results by means of a test of twenty minutes 
duration. 

Report of Steam Engine Test 
Item. 

1. Date, November 15, 1915. 

2. Kind of Engine. Weston high speed automatic noncon- 

densing single cylinder and simple valve. 

3. Dimensions 10" X 13" Piston Rod 1%". 

4. Rated horsepower 75 I.H.P. 

5. Horsepower constant, head end 0.002578 

Horsepower constant, crank end 0.00250 

6. Atmospheric pressure, in mercury ' 28.768 

Atmospheric pressure, lbs. per sq. in 14.12 

7. Length of brake arm 5 ft. 

8. Brake constant 0.000952 

9. Duration of test 20 min. 



STEAM ENGINE TESTING 127 

Item. • 

10. Average R.P.M 243 

11. Average steam line pressure, lbs. gage 116 

Average steam line pressure, lbs. absolute 130.12 

12. Average M.E.P. from indicator diagrams H.E 62.83 

Average M.E.P. from indicator diagrams C.E 61.51 

13. Total weight condensed steam, lbs 1113.5 

14. Weight of condensed steam, lbs. per hour 3340.5 

15. Quality of steam, per cent 98.5 

16. Dry steam supplied per hour, lbs 3303.7 

17. Brake load, net, lbs 306 

18. Brake horsepower 70.78 

19. Indicated horsepower, head end 39.36 

20. Indicated horsepower crank end 37.37 

21. Indicated horsepower, total ; 76.73 

22. Friction horsepower 5.95 

23. Mechanical efficiency, per cent 92.2 

24. Dry steam per I.H.P. hr., lbs 43.06 

25. Dry steam per B.H.P. hr., lbs 46.8 

26. Exhaust pressure, lbs. absolute from indicator diagrams. 18 

27. Thermal efficiency, on I.H.P. per cent 5.916 

28. B.t.u. supplied per I.H.P. hr. above exhaust pressure. . . . 43,015 

Calculating Results. — Item 5. — The formula for calculating 
the indicated horsepower is 

IHP = ^^^ 
33,000 

which gives the indicated horsepower developed on one side of 
the piston if n in the formula is the revolutions per minute. For 
any particular engine the length of stroke I is a constant quantity; 
the area of the piston, a, is constant; and the quantity 33,000 

is constant. Therefore ^r. r.r.r. will be a constant quantity and 

it is this part of the horsepower formula that is called the ''horse- 
power constant" or 

H.P. constant = ^^ ^^^ 

The reason for calculating the horsepower constant separately 
instead of calculating the horsepower directly is that it saves 
considerable time when the horsepower must be calculated a 
large number of times from indicator diagrams. The mean 
effective pressure from these diagrams will vary sHghtly, as will 
also the number of revolutions per minute. If the horsepower 
constant is calculated separately, this part of the calculations 



128 STEAM ENGINES 

need be done only once because then the horsepower may be 
calculated for any M.E.P. and R.P.M. by merely multiplying 
the engine constant by the M.E.P. and the R.P.M. It is 
necessary to calculate the horsepower constant for each end of 
the cylinder because the area of the piston is larger on the head 
end than on the crank end, since the area of the piston rod cuts 
off some of the area of the piston. 

13 

For this engine the length of stroke is 13 in. or y^ ft. and the 

area of the piston is 10^ X .7854 = 78.54 sq. in. The head end 
horsepower constant is therefore 

u r) X ^ 1 1 1 ^« 13 X 78.54 ^^^^^^ 

H.P. constant, head end = 33^ = 12^^3,000 = '^^^^^^ 

In calculating the horsepower constant for the crank end it is 
necessary to deduct the area of the piston rod from the area of 
the piston. The area of the piston rod is 

1.752 X .7854 = 2.405 sq. in. 
As the area of the piston is 78.54 sq. in., the net effective area is 

78.54 - 2.405 = 76.13 sq. in. 
Therefore the crank end horsepower constant is 

H.P. constant, crank end = ~^^ = ^f~J^ = -0025 

Item 6. — The atmospheric pressure is read on a barometer and 
is expressed in inches of mercury. To express the atmospheric 
pressure in pounds per square inch it is necessary to multiply by 
.4908 or 

Atmospheric pressure, lbs. per sq. in. = 28.768 X .4908 
= 14.12. 

Item 8. — The brake constant is a constant quantity similar to 
the ''horsepower constant" but relating to the Prony brake and 
it is used for lessening the work of calculating the brake horse- 
power. The formula used for calculating the brake horsepower 
with a Prony brake is 

_2^RWN 
33,000 
In which 

R = the length of the brake arm in feet 
W = the net weight in pounds registered by the brake 
and 

N = the number of revolutions per minute. 



STEAM ENGINE TESTING 129 

For any given engine test the weight registered by the brake and 
the R.P.M. may vary but the other parts of the formula will 
remain constant. Therefore 

Brake Constant = ooTaqTi = ^"^ non ~ -000952. 

Item 14. — Item 13 gives the weight of condensed steam in 
20 minutes, therefore the weight of condensed steam in one hour 
will be 

Item 13 X 2?) or 

* fKCi 

1113.5 X ~ = 3340.5 

Item 16. — The number of heat units in one pound of steam 
supplied to the engine is found by the formula 

qL + h 
in which q = the quality of the steam 

L = the latent heat per pound 
and h = the heat of the liquid per pound. 
Substituting in this formula from the steam table the values for 
130.12 lb. per sq. in. absolute 

L = 872.2 and h = 319.4 

or B.t.u. per lb. = qL -{- h = .985 X 872.2 + 319.4 = 1178.5 
The total amount of heat supplied to the engine in one hour 
equals the weight of steam actually supplied per hour multiplied 
by 1178.5 or 

3340.5 X 1178.5 = 3936779.25 B.t.u. 
If the steam had been perfectly dry, it would have contained 
L -\- h heat units per pound, or 

L + h = 872.2 + 319.4 = 1191.6 B.t.u. per pound 

Therefore the equivalent weight of dry steam supplied per hour is 

3936779.25 -^ 1191.6 = 3303.7 lbs. per hour. 

Item 18. — As explained before the brake horsepower is found 
by multiplying together the net brake load in pounds (Item 17), 
the R.P.M. (Item 10), and the brake constant (Item 8) or 
B.H.P. = 306 X 243 X .000952 = 70.78 

Item 19. — In a similar manner the indicated horsepower, 
head end, is the product of the M.E.P., head end (Item 12), the 
R.P.M. (Item 10) and the head end horsepower constant (Item 
5) or 



130 STEAM ENGINES 

Head end I.H.P. = 62.83 X 243 X .002578 = 39.36 
Crank end I.H.P. = 61.51 X 243 X .0025 = 37.37 
Total I.H.P. = 39.36 + 37.37 = 76.73 

Item 22. — The friction horsepower is equal to the difference 
between the indicated horsepower and the brake horsepower or 

Item 22 = Item 21 - Item 18 
Friction H.P. = 76.73 - 70.78 = 5^ 

Item 23. — The mechanical efficiency is equal to the brake 
horsepower divided by the indicated horsepower or 

1.T 1 • 1 ^ . Item 18 70.78 ^^^ ^^ ^ 
Mechanical efficiency = j, ^ = ^^ „^ = .922 or 92.2 per 

cent. 

Item 24. — The dry steam per I.H.P. per hour is equal to the 
dry steam per hour (Item 16) divided by the indicated horse- 
power (Item 21) or 

3303.7 
Dry steam per I.H.P. hr. = ^^ „ ~ = 43.06 

Item 25. — In a similar manner the dry steam per B.H.P. hr. 

is equal to the dry steam per hour divided by the B.H.P. or 

3303.7 
Dry steam per B.H.P. hr. = ~^ ^ ' = 46.8 

Item 27. — The thermal efficiency of the engine, based on the 

I.H.P. may be calculated by the formula 

^^ . I.H.P. X 42.42 

Efficiency. = ^j^^-^^^rj;) 

in which I.H.P. is the indicated horsepower (Item 21) 

W is the actual weight of steam supplied to the engine 
per minute or Item 14 -^ 60 

q is the quality of the steam supplied to the engine, 
Item 15 

L is the latent heat per pound of the steam supplied to 
the engine 

h is the heat of the liquid per pound of the steam sup- 
plied to the engine 

hi is the heat of the liquid per pound of steam at exhaust 
pressure 
In this case I.H.P. = 76.73 

W = ^^^^ = 55.675 
bO 



STEAM ENGINE TESTING 131 

q = .985 

L = 872.2 

h = 319.4 

hi = 190.5 



Therefore 



76.73 X 42.42 

iLtliciency 55 ^75 ( ggg >< §72.2 + 319.4 - 190.5) 

_ 3254.88 ^ 3254.88 

~ 55.675 X 988 55006.9 *^ 

or 5.916 per cent. 

Item 28. — In calculating Item 27 it was seen that the number 
of heat units in one pound of steam above exhaust pressure as 
actually supplied was 988. Since there was supplied to the en- 
gine 3340.5 lbs. of steam per hour, the number of heat units 
supplied per hour above exhaust pressure was 

988 X 3340.5 = 3,300,414 
and as the I.H.P. was 76.73 the number of heat units supplied 
per I.H.P. per hour above exhaust pressure was 

3300414 



76.73 



= 43015 



Duty of Pumps. — The performance of pumping engines is 
usually stated in terms of the number of foot-pounds of work 
performed by the water piston of the pump per thousand pounds 
of dry steam, or per million B.t.u. consumed by the engine. 
The performance, stated in this way, is called the duty of the 
pumping engine; thus, 

j^ . _ Foot-pounds of work done 

Weight of dry steam used 

-^ , Foot-pounds of work done , ^^^ ^^^ 

Duty = „, . /l . . 1 -^ X 1,000,000 

Weight 01 dry steam used 

In using the above formulas it is to be understood that the foot- 
pounds of work and the weight of steam or number of B.t.u. 
consumed are to be taken for the same periods of time. 

Example. — A compound pump uses 80 pounds of steam per I.H.P. per 
hour and develops 48 I.H.P. The pump receives steam at an absolute 
pressure of 135 pounds per sq. in. and exhausts against an absolute pressure 
of 17 lbs. per sq. in. The quality of the steam delivered to the pump is 97 
per cent. The capacity of the pump is 400 gallons per minute and pumps 
against a pressure of 175 pounds per sq. in. Calculate the duty of the 
pump, on the dry steam and on the heat unit bases. 
13 



132 STEAM ENGINES 

Solution. — 175 lb. per sq. in. pressure is equivalent to 

175 X 2.3 = 402.5 feet head 
400 gallons is equivalent to 

'400 X 8.33 = 3332 pounds 
Work done by pump per minute 

= 3332 X 402.5 = 1,341,130 foot-pounds 
Work done by pump per hour 

= 1,341,130 X 60 = 80,467,800 foot-pounds 
Heat units in one pound of wet steam above 32° 

= (.97 X 869.9 + 321.7) = 1165.5 B.t.u. 
Heat units in one pound of dry steam above 32° 

= 1191.6 B.t.u. 
Weight of wet steam used per hour 

= 80 X 48 = 3840 pounds 

Equivalent weight of dry steam used 

1165 5 
= 3840 X TT^a = 3755 pounds 
1191.6 

Duty = ^^'g^^^^^^ X 1000 = 21,423,800 foot-pounds 

Heat units in one pound of steam above heat of liquid at exhaust pressure 

= 1165.5 - 187.5 = 978.0. 
Heat supplied to pump per hour above heat of liquid at exhaust pressure 

= 978.0 X 3840 = 3,754,520 B.t.u. 

Duty = ^^'j^Jf20 ^ 1;000,000 = 21,432,247 foot-pounds. 



CHAPTER X 
THE SLIDE VALVE 

Steam and Exhaust Lap. — The valves controlHng the distribu- 
tion of steam to the cyhnder are the most important parts of a 
steam engine, because both the smooth running of the engine and 
the economical use of steam depend largely upon them. 

Of the different kinds of valves used on engines, the ordinary 
slide valve is the most important. The slide valve combines in 
one valve the office of admission and exhaust for both ends of the 
cylinder ; it is used on a greater variety of engines than any other 
form of valve; and the principles underlying the operation of the 
slide valve are also the principles underlying the operation of 
other types of valves. For these reasons a thorough study of 
the slide valve will be made before considering other types. 

The operation of the slide valve is described in Chapter 1, 
and the valve and its mechanism are illustrated in Figs. 2 and 4. 
The eccentric which moves the valve backward and forward 
gives it the same motion as would a crank, and in fact it is equiva- 
lent in all respects to a crank having a length equal to the dis- 
tance from the center of the shaft to the center of the eccentric. 
This distance is called the eccentricity. 

The position of a line connecting the center of the shaft with 
the center of the eccentric also represents the position of the 
eccentric or the position of the equivalent crank. The distance 
which the valve travels in going from one end of its stroke to 
the other is called the valve travel. If the motion of the eccen- 
tric is communicated directly to the valve, the valve travel will 
equal twice the eccentricity. 

When a valve is at the middle point of its travel, in which 
position the eccentric will be vertical to the center line of the 
engine, the valve is in mid-position. This position of a valve is 
used as a reference point from which the parts of the valve and 
also its different positions are measured. The cross section 
of a slide valve in its mid-position is shown in Fig. 74. When 
it is in this position, the length from the outer edges of the ports 
14 133 



134 



STEAM ENGINES 



P and P to the outer edges A and A of the valve is called the 
outside lap. The outside lap is shown by the dimensions and 
0. It is not necessary that the outside lap at one end of the valve 
be equal to that at the other end, and, in fact, they are usually 
unequal, as will be explained later. The distance from the inside 
edges of the ports P and P to the inner edges B and B of the valve 
is called the inside lap. The inside lap is shown by the dimen- 
sions I and I. The inside laps of a valve are usually unequal. 
The width of the ports, or the distance /, is usually the same for 
both ends of the cylinder. 

When steam is admitted past the outer edge of the valve, the 
outside lap is usually called the steam lap and the inside lap 
the exhaust lap. When steam is admitted from inside the valve 




Fig. 74. 

and exhausted past the outer edge, as is sometimes done, the 
inside lap is called the steam lap and the outside lap is called 
the exhaust lap. The steam lap is usually much greater than the 
exhaust lap, hence a valve designed for outside admission will 
not distribute the steam properly, if it is used for inside admission. 
Unless otherwise mentioned the outside lap will be considered 
as the steam lap as this is the more usual arrangement. 

Valve Without Laps. — A form of slide valve often used on 
small direct acting steam pumps is shown in Fig. 75. It will be 
observed that this valve differs from the one shown in Fig. 74 
in having neither steam nor exhaust laps, the width of the valve 
being just equal to the width of the ports. In the position shown 
in Fig. 75 the valve is in its mid-position and, since it has neither 
steam nor exhaust lap, it will open the port to admission on one 



THE SLIDE VALVE 



135 



end and to exhaust on the other if the valve moves ever so Httle 
to either side of its mid-position. If this kind of sUde valve were 
to be used on a steam engine it would have to be set so as to be 
in its mid-position when the piston was at the end of its stroke. 
This would require that the eccentric be placed 90° from the 
crank as shown at the right of Fig. 75, where Oc represents the 
position of the crank and Oe the position of the eccentric. With 
the valve in its mid-position and the piston at the end of its 
stroke, steam will begin to be admitted to the cylinder as soon as 
the piston starts forward and the valve will remain open until 
the piston has reached the end of its stroke, thus giving admission 
throughout the entire stroke. At the end of the stroke the valve 
will close the admission and open the exhaust, and exhaust will 
occur throughout the entire return stroke of thefpiston. It 





Fig. 75. ' 

should be noted that the eccentric may be set either in the posi- 
tion Oe, Fig. 75, or in the position Od, and the engine will run 
either in a clockwise or counterclockwise direction of rotation, 
depending upon the direction in which it starts. The only 
requirement in setting the eccentric for a valve with no laps 
is that it must be placed 90° from the crank. 

An indicator diagram taken from an engine fitted with a valve 
having neither steam nor exhaust lap will be simply a rectangle, as 
shown in Fig. 76. This diagram shows that steam is admitted 
to the cylinder during the entire stroke, being released at the 
end of the stroke without having expanded. Exhaust also occurs 
during the entire return stroke and none of the steam is com- 
pressed near the end of the stroke. 

A valve without laps is suitable for small direct acting pumps 
because the load on these pumps is a constant water pressure and 
the steam pressure on the piston must therefore be constant 
throughout the entire stroke in order to overcome the constant 



136 



STEAM ENGINES 



water pressure. Moreover, the constant load on the piston 
serves to bring it to rest at the end of the stroke without shock 
at the low speeds at which such pumps are usually run, hence 
no compression is necessary. A valve of this kind would not, 
however, be suitable for a steam engine because it would be 
uneconomical in the use of steam since the expansive force of the 
steam would not be used, and also because the high speed of the 
engine makes compression necessary if the engine is to run 
smoothly. It is not necessary to admit steam to the cylinder 
of an engine throughout the entire stroke because the engine 
has a flywheel which stores up energy in the first part of the 
stroke, giving it out again in the last part of the stroke when the 



Admission 



Exhaust 



Fig. 76. 



steam pressure is small and, by this means, causing the engine 
shaft to rotate at a uniform speed. 

Valves With Lap. — A valve that has steam lap will keep the 
port closed against the admission of steam until the valve has 
moved from its mid-position a distance equal to the steam lap. 
The valve shown in Fig. 77 is an outside admission valve and it 
has moved to the right of its mid-position a distance equal to the 
steam lap. In this position the valve is just on the point of 
admitting steam to the head end of the cylinder. If the valve 
now moves to the right steam will be admitted to the cylinder 
and admission will continue until the valve moves to the left 
and returns to the position shown in Fig. 77 when the port will 
be closed. The port will then remain closed and the steam will 
expand until the valve moves far enough to the left to bring 
the inner edge of the valve in line with the inner edge of the port. 



THE SLIDE VALVE 



137 



A further movement of the valve to the left uncovers the port 
for exhaust, which continues until the valve, in moving to the 
right on its forward stroke, reaches a position where the inner 
edge of the valve is again in line with the inner edge of the port. 
As the valve continues its movement to the right the port remains 
closed and the steam in the cylinder is compressed. Compres- 
sion continues until the valve, still moving towards the right, 
reaches a position in which the outer edge of the valve is in line 
with the outer edge of the port, when admission again occurs. 
It is thus seen that the purpose in having steam and exhaust laps 
is to permit the steam to be expanded and compressed. 

Position of Crank and Eccentric. — In Fig. 77 the position of the 
valve is such that steam is about to be admitted to the head end 



Pn 





Fig. 77. 

of the cylinder and the piston is shown at the beginning of its 
forward stroke. If the valve is connected to the eccentric with- 
out the use of a rocker arm, the corresponding positions of the 
crank and eccentric will be as shown in the diagram at the right 
of Fig. 77, in which represents the center of the shaft, OC the 
position of the crank, and OE the position of the eccentric. 
When the valve is in its mid-position the eccentric is vertical, 
in the position of the line OA, but in the position shown at OE 
it has been moved around on the shaft in a clockwise direction 
enough to move the valve through a distance equal to its steam 
lap, which is equal to the distance OL on the diagram. The 
distance OL on the diagram represents the displacement of the 
valve from its mid-position. 

With the positions of crank and eccentric as shown in Fig. 77 
if the shaft turns in a clockwise direction, as shown by the arrow, 
the piston will move forward and the valve will move to the 



138 STEAM ENGINES 

right, admitting the steam behind the piston. This will push 
the piston forward and cause the engine to run. If, however, 
an attempt is made to start the engine by turning the shaft in a 
direction opposite to that indicated by the arrow, or in a counter- 
clockwise direction, the valve will be moved to the left and prevent 
the admission of steam. It is seen then that, with the positions 
of crank and eccentric as shown, the engine can run only in a 
clockwise direction. // there is no rocker arm, the direction of ro- 
tation of the engine will be such that the crank follows the eccentric. 
If it was desired to have the engine under consideration run in a 
counterclockwise direction the eccentric would have to be set 
in the position OF, the crank being at OC. 

Valve. Rod 




PIVOT 



, E ccentric Rod 

Q ) =:t 

Fig. 78. 

Sometimes the shape of the engine is such that the valve and 
eccentric are not jn line with each other and the motion is trans- 
mitted from the eccentric to the valve through a rocker arm as 
shown in Fig. 78. If the valve rod and eccentric rod are both 
on the same side of the pivot the motion of the valve will be in 
the same direction as if the valve rod was connected directly 
to the eccentric rod, but if they are connected on opposite sides 
of the pivot, as shown in Fig. 78, the valve will move in the oppo- 
site direction to that in which it would move if directly connected. 
Since a rocker arm pivoted between valve and eccentric rods 
reverses the motion of the valve, the position of the eccentric 
with respect to that of the crank will be as indicated by the line 
OF of the diagram at the right of Fig. 77 for a clockwise direction 
of rotation. In other words, the use of such a rocker arm requires 
that the eccentric be set to follow the crank. 

Lead. — In Fig. 77 the valve is set to admit steam to the cylinder 



THE SLIDE VALVE 139 

just at the beginning of the stroke. If a valve is set in this way 
there will be considerable drop in steam pressure during admis- 
sion, or ''wire drawing." The effect of this on the indicator 
diagram is shown in Fig. 79 in which the reduction in the admis- 
sion pressure is indicated by the drop in the admission line. 
Since this reduces the area of the diagram it shows that the power 
of the engine is reduced. It also reduces the efficiency of the 
engine because the range of pressure through which the steam 
may be expanded is reduced. 

An inspection of the diagram at the right of Fig. 77 will show 
that, if the shaft rotates with a uniform speed, the valve travels 
fastest when it reaches mid-position and the piston travels fastest 

^ BoiuER Pressure 



Fig. 79. 

when it reaches mid-stroke, the speed of each increasing during 
the first part of its stroke and decreasing during the last part. 
This diagram also shows that, at the beginning of the piston 
stroke, the speed of the valve is decreasing and the speed of the 
piston is increasing. The result is, that if the valve is set to 
open just at the beginning of the piston stroke, steam cannot 
flow into the cylinder through the narrow opening fast enough 
to maintain full pressure behind the piston, hence the pressure 
in the cylinder drops. 

In order to prevent excessive drop in pressure during admission 
the valve is set so as to open slightly before the end of the exhaust 
stroke thus insuring enough port opening to allow free admission 
when the piston starts forward. The amount which the port is 
open for admission when the piston is at the end of its stroke is 
called the lead of the valve. The amount of lead which a valve 
should have depends upon the size of the cylinder and speed of 
the engine, being larger for high speeds and large cylinders and 
smaller for slow speeds and small cylinders. 



140 



STEAM ENGINES 



The cushioning of the piston depends to a certain extent upon 
the lead as a large lead gives an early opening of the valve with 
an early admission of high pressure steam against which the 
piston must advance at the end of the exhaust stroke. 

A valve set with lead is shown in Fig. 80, the relative positions 
of the crank and eccentric being shown in the diagram to the 
right of the figure. In this illustration the piston is at the head 
end of its stroke and the valve is open an amount I for the admis- 
sion of steam to the head end. The distance I, in this case is the 
lead. It will be observed that, in the position shown, the valve 
has moved to the right of its mid-position a distance equal to 
the steam lap plus the lead. 

In the diagram at the right of Fig. 80 the crank OC is shown on 
dead center to correspond with the position of the piston. The 



on 




2 c 




Fig. 80. 

line OE shows the position of the eccentric and the distance OL 
shows the displacement of the valve from its mid-position. If 
the valve was set without lead the position of the eccentric would 
be OE^, the distance OL' being equal to the steam lap. Since 
the distance OL is equal to the steam lap plus the lead, the 
distance LX is equal to the lead of the valve. It will be observed 
that when the valve has lead, the eccentric must be moved 
around on the shaft far enough to displace the valve from its 
mid-position a distance equal to the steam lap plus the lead when 
the crank is on center. 

Angle of Advance. — For an outside admission valve connected 
directly to the eccentric, the eccentric must be set so that it 
leads the crank. When the crank is on center the valve must 
also be displaced to the right of its mid-position a distance equal 
to the steam lap plus the lead. These two conditions determine 
the position of the eccentric with respect to that of the crank. 



THE SLIDE VALVE 



141 



If the eccentric was in the position OA, Fig. 80, when the crank 
was on center, the valve would be in its mid-position, hence the 
eccentric must be moved forward through the angle AOE in 
order to displace the valve from its mid-position a distance equal 
to the steam lap plus the lead. This makes the angle between 
the crank and eccentric, which is called the crank angle, greater 
than 90°. The angle AOE is called the angle of advance and it 
is the angle, in excess of 90°, between the crank and eccentric. 
The angle of advance is usually about 20° to 30° but its amount 
will depend upon the steam lap and the lead which it is desirable 
to give the valve. The angle of advance is important in valve 
setting, as will be shown later, because it is the only thing about 
the valve mechanism besides the length of the valve rod which 
may be adjusted. 




3^ 



- '^- - 









Fig. 81. 



Inside Admission Valve. — Many valves, especially of the piston 
type, are designed to admit steam from the inside and exhaust 
past the outer edge of the valve. In this case the inside lap is 
the steam lap and the outside lap is the exhaust lap. A slide 
valve arranged for inside admission is illustrated in Fig. 81. As 
shown here the piston is at the head end of its stroke and steam 
is being admitted to the head end of the cylinder. The valve 
is therefore displaced to the left of its mid-position a distance 
equal to the steam lap plus the lead, and the eccentric must be 
on the left-hand side of the vertical line A 5 in the diagram to the 
right of Fig. 81. In order for the engine to run the valve must 
move to the left when the piston starts on its forward stroke. 

If this valve is connected directly to the eccentric, without a 
rocker arm, and the rotation is to be clockwise the eccentric must 
be in the position OE when the crank is in the position OC, since 



142 STEAM ENGINES 

in this position the valve will be moved to the left and be opened 
a greater distance when the crank moves in a clockwise direction. 
If the crank should move in a counterclockwise direction the valve 
would close and cut off the supply of steam when the piston started 
forward. Suppose the eccentric was placed at OE' when the 
crank is at OC. The valve would then be in the position shown 
in Fig. 81 but a clockwise rotation of the shaft would move the 
valve to the right and close the port. However, if the shaft 
rotates in a counterclockwise direction the valve would move 
to the left and open the port further when the piston starts 
forward. It may be stated, then, that for an inside admission 
valve without a rocker arm the eccentric should he set to follow the 
crank. Since a rocker arm reverses the direction of motion of 
the valve, the presence of a rocker arm requires that the eccentric 
be set to lead the crank, that is, in Fig. 81 if there were a rocker 
arm between the eccentric and the valve, a clockwise direction of 
rotation would require that the eccentric be set at OE^ and a 
counterclockwise rotation would require that it be set at OE. 
In Fig. 81 the angle of advance is BOE and it is negative, since 
the crank angle is less than 90°. The angle of advance for a 
valve with inside admission does not differ in amount from that 
for a valve with outside admission, since its amount depends 
only upon the steam lap and the lead, but it does differ in position. 



CHAPTER XI 
THE VALVE DIAGRAM 

Valve Displacement. — Since an eccentric is equivalent to a 
crank, it may be represented as a crank having a length equal 
to the eccentricity. In Fig. 82 the valve V is connected by the 
eccentric rod E to the eccentric OC, the eccentric being repre- 
sented here by a crank having a length OC equal to the distance 
from the center of the shaft to the center of the eccentric. The 
circle ABCF represents the path followed by the center of the 
eccentric as the shaft rotates — and its diameter AB shows the 
length of the valve travel, which is twice the eccentricity OC. 

When the center of the eccentric C is at the point A, the valve 
is at the extreme left of its travel, when C is at the point B the 



C,JH 




Fig. 82. 



valve is at the extreme right of its travel, and when C is at // 
the valve is in its mid-position. As C passes through the half 
circle AHB the valve moves through a distance equal to that 
from A to B, and as C passes through the half circle BKA the 
valve moves through a distance equal to that from B to A. The 
diameter AB represents the valve travel to the same scale that 
OC represents the eccentricity, and the position of the valve 
at any time during its travel may be located on the diameter AB. 
It will be observed that the eccentric rod is long in comparison 
with the valve travel, and that during a revolution of the shaft 
the eccentric rod never makes a large angle with the center line 
BL. When this is the case the position of the valve may be 
located for any position of the eccentric by simply projecting 
vertically to the diameter AB, the point representing the center 
of the eccentric. Thus, when the eccentric is in the position 

143 



144 



STEAM ENGINES 



ii. , 



-4 



OC the valve will be at the point D in its travel, the point D being 

found by drawing CD at right angles 
^, '"I"""" "J to AB. The distance OD is the dis- 

placement of the valve from its mid- 
position. When the eccentric is at 
OF, G represents the position of the 
valve, and OG its displacement from 
mid-position. It will be observed 
that if i^ is exactly opposite C, the 
valve displacement OG when the 
eccentric is at OF is the same as its 
displacement OD when the eccentric 
is at OC. 

Piston Position. — The same kind of 
diagram as shown in Fig. 82 may be 
used to represent the travel of the 
piston, but the method of locating 
the position of the piston differs on 
account of the fact that the connect- 
ing rod is usually shorter when com- 
pared to the length of the piston travel 
and that at certain parts of the revo- 
lution of the crank, the connecting 
rod makes a considerable angle with 
the center line of the engine. 

In Fig. 83 the circle AHBK repre- 
sents the path followed by the center 
of the crank pin C during a revolution 
of the shaft whose center is at 0. OC 
is the length of the crank and the 
diameter A 5 of the crank circle rep- 
resents the piston stroke. When the 
crank is in any position as OC, the 
corresponding position of the piston 
may be located by taking a radius 
equal to the length of the connecting 
rod MC, and with M as a center 
drawing an arc CD through C until 
it strikes the diameter AB at the 
point D. The point D will represent 
the position of the piston for the posi- 



00 



o 



I 



z 
o 
h 
52 
a 

1. 



THE VALVE DIAGRAM 145 

tion OC of the crank and the distance AD shows how far the 
piston is from the end of its stroke when the crank is at OC. 
The apparent position of the piston is at R, obtained by pro- 
jecting the point C vertically on the diameter AB as was done 
for the valve position in Fig. 82. The actual position D of the 
piston is displaced from its apparent position R by the distance 
RD and this is due to the comparatively large angle which the 
connecting rod makes with the center line LB of the engine. 
When the crank is in the position OF, exactly opposite OC, the 
actual position of the piston is at G while its apparent position 
is at S. This effect is due to the ajigularity of the connecting rod 
and its amount depends upon the length of the connecting rod 



Fig. 84. 

as compared with the length of piston stroke. In Corliss engines 
the length of connecting rod is so great, as compared with the 
piston stroke, that the angularity of the connecting rod seldom 
need be taken into account. Other types of engines, on the 
other hand, have comparatively short connecting rods and the 
angularity must be considered in locating the position of the 
piston for cut-off, release, and compression, and sometimes also 
for admission, which occurs nearer the end of the stroke than any 
of the other events and hence is not so much affected by the 
angularity of the connecting rod. 

Position of Crank and Eccentric. — In Fig. 82 it is shown that 
the position of the eccentric and the valve displacement may be 
represented on a circle having a radius equal to the eccentricity 



146 STEAM ENGINES 

and in Fig. 83 it is shown that the position of the crank and piston 
may be represented on a circle having a radius equal to the length 
of the crank. The corresponding positions of crank and eccentric 
may therefore be represented by two circles drawn about the 
same center, one for the crank and one for the eccentric, as shown 
in Fig. 84. 

In most engines the eccentricity is small as compared with the 
length of the crank, hence if both the crank circle and the eccen- 
tric circle are drawn to the same scale, one will be large and the 
other small, and it may happen that the eccentric circle will be so 
small as to cause difficulty in making measurements tupon it. 
In Fig. 84 the length of the crank OC is 12 inches, giving a stroke 

of 24 inches, and the eccentricity OE is 2 

inches, giving a valve travel of 4 inches, 
these being common proportions between 
valve travel and stroke. In order to draw 
the corresponding positions of crank and 
eccentric upon these circles, the position 
of the crank OC is first drawn. Then by 
laying off the crank angle COE, the eccentric 
OE may be drawn. 
It will be seen from Fig. 84 that if both the crank circle and. 
the eccentric circle be drawn to the same scale it will be difficult 
to measure valve displacements accurately on account of the 
small size of the eccentric circle. In order to avoid this difficulty 
the eccentric circle may be drawn the same size as the crank 
circle, thus making a single circle serve for both crank circle and 
eccentric circle, as in Fig. 85. If this is done the diameter of the 
circle will represent the piston stroke to one scale and the valve 
travel to a different scale. Thus, in Fig. 85 the diameter AB 
represents a piston stroke of 24 inches and it also represents a 
valve travel of 4 inches. For the crank position OC the piston 
is at a distance AD from the end of its stroke and this distance 
is 3.6 inches on the scale by which the diameter represents 24 
inches. The valve displacement corresponding to the crank 
position OC is shown at OF and this distance is 1.6 inches on the 
scale by which the diameter represents 4 inches. 

Valve Diagram. — A diagram may be drawn which shows the 
valve displacement for all positions of the crank. Such a dia- 
gram is called a valve diagram. A valve diagram is useful in 
setting the valve because it shows at a glance the effects of any 



THE VALVE DIAGRAM 147 

changes which may be made in the valve or eccentric and thus 
tells what changes to make in order to accomplish desired results. 
The diagram shown in Fig. 85 might be used as a valve diagram 
since from it could be obtained the valve displacement for any 
position of the crank, but it is not convenient to use this dia- 
gram because the crank angle must be laid off for each new 
position of the crank at which it is desired to measure the valve 
displacement. 

The most common form of valve diagram is called the Zeuner 
diagram, after the name of its inventor. The Zeuner valve 
diagram is drawn as follows: The circle ACB in Fig. 86 is 
drawn so that its diameter represents the piston stroke to one 
scale and the valve travel to another scale. Imagine the eccen- 
tric to be in the position OB ; the crank will e 
then be in the position OC and the angle y/\^\ ^\ 
COB will represent the crank angle. With Tl^x ! ) \ 
the eccentric in the position OB the valve .t V^^\l/ 1 ^ 
will be at the extreme right of its travel \ | y 
and its displacement from mid-position \ | / 
will be a maximum, being equal to the ^^"^^^i,^--^ 
eccentricity. Using the line OC, which 
represents the crank, as a diameter, draw 

the small circle OPC. This is called the valve circle. Any radial 
line drawn from to represent any position of the crank, and 
cutting the valve circle, will show the valve displacement from 
mid-position by the length which the valve circle cuts off on 
it. Thus, when the crank is in the position OC, the valve dis- 
placement is equal to the length OC. When the crank is in any 
other position, as at ON the valve displacement is OP, being 
always the length which the valve circle cuts off on the radial 
line which represents the crank position. 

It is to be observed that in Fig. 86 the angle COE represents 
the angle of advance since the angle of advance is equal to the 
crank angle minus 90°. In this case the angle COB is the crank 
angle and taking away 90°, or the angle EOB, leaves the angle 
COE as the angle of advance. 

In the preceding chapter it was shown that the valve displace- 
ment at admission and cut-off, is equal to the steam lap. There- 
fore, if the arc of a circle be drawn with O, Fig. 87, as a center and 
a radius OL equal to the steam lap, it will cut the valve circle at 
the points H and J and a line OD drawn through and H will 



148 



STEAM ENGINES 



represent the position of the crank at admission, when the valve 
displacement is equal to OH, the steam lap. Also, a line OG 
drawn through and J will represent the position of the crank at 
cut-off when the valve displacement is again equal to the steam 
lap, OJ. The arc HLJ is called the steam lap circle. Fig. 87 
is drawn in the same way as Fig. 86, but is made separate in 
order to avoid confusion. 

The lead of a valve is the amount of port opening when the 
crank is on dead center. The port opening is equal to the valve 
displacement minus the steam lap. In Fig. 87, OA represents 
the position of the crank when on dead^center. OM shows the 




valve displacement for this position of the crank and OL the 
steam lap, hence LM represents the lead of the valve. The 
amount of port opening is always the distance between the lap 
circle, and the valve circle, as shown by the shaded area in Fig. 
87. 

It should be observed that the diagram in Fig. 87 shows only 
valve displacements to the right of mid-position and for this 
reason the valve circle is marked R. In order to show valve 
displacements to the left of mid-position another valve circle 
U must be drawn opposite the one marked R, as shown in Fig. 
88. The valve circle marked U for showing displacements to the 
left must be the same size as the other one and must be drawn 
on the same line 00 extended through the crank circle. Since 
release and compression occur when the valve displacement is 



THE VALVE DIAGRAM 



149 



equal to the exhaust lap, the lap circle TS must be drawn with a 
radius equal to the exhaust lap. The four positions of the crank, 
at admission, cut-off, release, and compression may now be 
drawn on the diagram as shown in Fig. 88. Since admission and 
cut-off, as shown here, take place when the valve is displaced to 
the right, and release and compression take place when the valve 
is displaced to the left, these events are all for one end of the 
cylinder only. In Fig. 88 the events shown are for the head end 



CUT -OFF 



ADMISSION 



COMPFie. 




RELEASE. 



-J B 



Head End Indicator Diagram 

Fig. 88. 

of the cylinder, if there is no rocker arm, the direction of rotation 
being clockwise, as indicated by the arrowhead. 

With the crank positions at admission, cut-off, release, and 
compression as shown in Fig. 88 the approximate shape of the 
indicator diagram that will be obtained from this valve setting 
may be shown if the admission and exhaust pressures are assumed. 
In order to draw this indicator diagram the crank positions are 
projected to the line AB with a radius equal to the length of the 
connecting rod (to the same scale that AB represents the piston 

15 



150 STEAM ENGINES 

stroke) . These points are then projected vertically to the admis- 
sion and exhaust lines below and the indicator diagram sketched 
in as shown, using smooth curves to connect cut-off and release, 
and also compression and admission. 

The effects of certain changes in the valve setting may readily 
be observed from the valve diagram in Fig. 88. For a given 
eccentric the eccentricity is a fixed quantity and cannot be 
changed. The steam and exhaust laps are parts of the valve and 
cannot be readily changed although they may be decreased 
slightly by chipping or filing off the end of the valve. This 
leaves only the angle of advance that may be changed readily, 
which may be done by shifting the eccentric around on the shaft 
either to increase the angle of advance or to decrease it. 

The angle COE in Fig. 88 represents the angle of advance. It 
will be observed from the valve diagram that if the angle of 
advance is made larger admission, cut-off, release, and compres- 
sion will all occur earlier in the stroke. If the angle of advance 
is made smaller all of these events will occur later in the stroke. 
It will be observed also that, other things being left unchanged, 
a large steam lap gives late admission and early cut-off and a large 
exhaust lap gives late release and early compression. The steam 
and exhaust laps might be increased by fastening a block to the 
outer or inner edges of the valve but this is not often done as it is 
inconvenient and not often necessary since the valve is designed 
with the proper laps. A smaller steam lap causes early admission 
and late cut-off and a small exhaust lap causes early release and 
late compression. As explained before, the steam and exhaust 
laps may be made slightly smaller by filing or chipping, but this 
is not often necessary. 

It will be further observed from the valve diagram that the 
angle which the crank turns through while the steam is being com- 
pressed in the cylinder, is the same as the angle turned through 
while the steam is expanding. This sets an important limitation 
upon the action of the slide valve because it does not permit 
so large a degree of expansion of the steam with economy as 
might otherwise be secured. In order to secure good economy 
an engine must expand the steam through a large range of pres- 
sure. The range of pressure through which steam may be expanded 
depends upon the point of cut-off; if cut-off is early the range 
of J pressure will be large, but if cut-off is late the range of pressure 
will be small. If it is attempted to secure ah early cut-off 



THE VALVE DIAGRAM 



151 



with a slide valve the point of compression will also be early. 
If the cut-off is made earlier than about half stroke, the gain from 
greater expansion will be more than counterbalanced by the loss 
from greater compression. This result may be seen from Fig. 
89 in which the full line diagram shows cut-off at seven-eighths 
of the stroke with the corresponding compression, and the dotted 
lines show the greater expansion for cut-off at half stroke with the 
corresponding compression. With the cut-off at half stroke the 
gain in economy from the greater expansion is counterbalanced 
by the loss of area from the indicator diagram by the earlier 
compression. The action of the slide valve described above ex- 
plains why an engine fitted with this type of valve is uneconom- 




FiG. 89. 

ical in the use of steam; that is, it cannot use the full expansive 
force of the steam as can engines which have separate valves for 
admission and exhaust. 

The valve diagram in Fig. 88 shows the events occurring in 
only one end of the cylinder, namely, the head end. The same 
diagram may also be used for showing events occurring in the 
crank end of the cylinder by drawing the crank end steam lap 
circle in the left (L') valve circle and the crank end exhaust lap 
circle in the right (R) valve circle. This has been done in Fig. 
90, the steam and exhaust lap circles for the head end being 
drawn with full lines and the steam and exhaust lap circles for 
the crank end being dotted. The crank positions for events in 
the head end are also drawn with full lines and those for the 
crank end with dotted lines. This makes a complete valve 
diagram which shows all of the actions occurring in both ends 
of the cylinder and shows also the effects produced by any changes 
in the valve or its setting. 



152 



STEAM ENGINES 



All of the valve diagrams shown up to the present time have 
been for a clockwise direction of rotation and for a direct con- 



CUT- OFF 
H END 

"C - END 



-■I- J 



/ \ / ^ 


^ 

\ 


1 
1 




.-^ ^""^ 1 


\ / ' 


\ 


V 1 




rTTSx ._. 


^^■^A=^=^ 


'/ 


//I 


\ / 


/ \ 



RELEASe 
H END 



Aoni3sior\) 

C. -END 



ADMISSION 
H. END -^ 

RELEASE 

c end' 



-i ^ 



COMPRESSION 
H END 
CUT- OFF_ 

C.ENO 



Fig. 90. 



nection between valve and eccentric. Fig. 91 shows how the 
valve diagram should be drawn for a counterclockwise direction 



COMPRESSION 
H. END 



RELEASE. 
C. END " 



ADMISSION 
H. END 




ADMISSION 
C.ENO 



RE.LEA3E. 
HENO 



COMPRESSION 
C. END 



of rotation. In this case the angle of advance COE is laid off 
to the right of the vertical line EF instead of to the left of it as 



THE VALVE DIAGRAM 153 

with clockwise rotation; and in all cases the valve circles are 
drawn on the line representing the position of the crank when the 
eccentric is on dead center. In Fig. 91 the crank will be at 
OC when the eccentric is at OA , hence the valve circles are drawn 
on the diameter through C and 0. Valve displacements to the 
right of mid-position are then measured on the bottom valve 
circle R and those to the left are measured on the top valve circle 
U . Admission and cut-off for the head end are therefore located 
by means of the bottom valve circle and release and compression 
for the head end by means of the top valve circle. Events for 
the crank end of the cylinder are located by means of the opposite 
valve circles from those for the head end events. 

The late cut-off generally employed with plain slide valve 
engines makes them uneconomical on account of not using much 




Fig. 91a. 

of the expansive force of the steam. With a late cut-off the 
terminal pressure, or pressure at the end of expansion will be 
high. When the exhaust port is opened, this high pressure 
remaining in the steam is wasted through the exhaust pipe. 

When such an engine is operated under a fairly constant load 
and not stopped at frequent intervals the steam consumption 
may sometimes be greatly reduced by redesigning its valve, 
changing its steam and exhaust taps, and re-setting the eccentric. 
Fig. 91a illustrates a set of indicator diagrams from an engine 
whose valve was redesigned to give an earlier cut-off. It can 
be seen from these diagrams that the steam distribution and 
economy is much better than is obtained with the ordinary slide 
valve cutting off at about three-quarters stroke. This valve was 
redesigned so as to give cut-off at about half stroke, release at 
about 90 per cent, of the stroke, and compression at about 72 
per cent, of the exhaust stroke. The angle of advance was also 
changed to about 50°. Making the cut-off earher increases the 



154 STEAM ENGINES 

probability that the engine will stop in a position after the valve 
has closed against admission which makes it impossible to start 
again until the flywheel has been turned, but if the engine is 
seldom stopped during the day, this is not likely to become a 
nuisance. 

Defects in the setting or adjustment of valves are readily de- 
tected by irregularities in the indicator diagram. A study of the 
valve diagram will show the causes of such defects in the valve set- 
ting and will suggest the remedy that should be applied. One of 
the most common defects in valve adjustment comes from shpping 
of the eccentric around on the shaft, the eccentric usually being 
fastened to the shaft by a single set screw. If the eccentric 
slips it is likely to do so against the direction of rotation, resulting 
in a decrease of the angle of advance. A decrease in the angle 




Fig. 92. 

of advance causes all of the events (admission, cut-off, release, and 
compression) to occur later. The effect on the indicator diagram 
is shown in Fig. 92, which was taken from an engine on which the 
eccentric had sHpped backward. The most noticeable defect 
due to the shpping backward of the eccentric is seen at release 
where the diagram will have a beak, as at 1, 2, showing that the 
valve does not open soon enough to allow the pressure in the 
cylinder to fall to the exhaust pressure before the piston starts on 
the return stroke. Shppage of the eccentric cannot be detected 
readily from the late cut-off because some engines have a much 
later cut-off than others. With a properly designed shde valve, 
however, late release is always accompanied by late cut-off. If 
the valve has proper lead before the eccentric slips backward, 
it will have too little lead afterwards. The effects of this on the 
admission line is shown by the rounding from 4 to 5 caused by the 
admission side of the valve opening too late, thus preventing the 
pressure in the cylinder from rising quickly and also causing 



THE VALVE DIAGRAM 



155 



wire-drawing. Small lead is not a good indication of a slipped 
eccentric because the valve may have been set originally with 
too little lead. The compression, in this case, is too late for a 
slide valve and gives another indication that the eccentric has 
slipped backward. 

If, by any means, the eccentric should become turned forward 




Fig. 93. 

on the shaft, the angle of advance will become larger, and all of 
the events will occur earlier. An indicator diagram showing the 
results of too great angle of advance is illustrated in Fig. 93. 
Too early release is shown at 3, 4, by the sharp toe of the diagram. 
Early admission, the result of too much lead, is indicated at 6. 




Fig. 93a. 

1, by the backward pointing beak. The cut-oif, 2, is evidently 
too early for a slide valve engine, although, in general, the 
position of the cut-off does not give a good indication of the 
valve setting. The early compression, 5, agrees with the early 
cut-off, as it should with a slide valve. 

Sometimes the valve becomes displaced on the valve stem. 



156 STEAM ENGINES 

resulting in the valve stem being either too long or too short 
for the valve setting desired. This will affect the events oc- 
curring in the two ends of the cylinder differently since the 
valve is moved bodily either to the right or to the left. If the 
valve stem is too long the steam lap on the head end is increased, 
which delays admission and hastens cut-off on the head end, as il- 
lustrated in Fig. 93a. The valve stem being too long also de- 
creases the exhaust lap on the head end, which hastens release and 
delays compression on the head end. The effects produced on the 
events of the crank end as illustrated in Fig. 936 are just opposite 




Fig. 936. 

from those produced on the head end. The steam lap on the 
crank end is reduced, which hastens admission and delays cut-off, 
while the crank end exhaust lap is increased, which delays release 
and hastens compression. 

The changes in events produced by shortening the valve stem 
will be the same as those produced by lengthening it, except that 
the changes produced in the head end events by lengthening the 
valve stem will be produced in crank end events if the valve stem 
is shortened, and the changes produced in the crank end events 
by lengthening will be produced in the head end if the valve stem 
is shortened. 



CHAPTER XII 
VALVE SETTING 

General Considerations. — In setting the valves of an engine the 
principal requirements are to secure an economical distribution 
of steam between the two ends of the cylinder and to secure 
smooth running of the engine. In order to obtain these results 
it is desirable to divide the work to be performed equally between 
the two ends of the cylinder. An unequal division of the work 
between the two ends of the cylinder throws unduly large strains 
upon the engine, is likely to cause variations in the speed, and 
causes the steam to be used under more unfavorable conditions 
in one end of the cylinder than in the other. An equal division 
of the work will be secured approximately when cut-off occurs 
at equal points of the admission strokes for the two ends of the 
cylinder, therefore, in setting engine valves it is desirable to secure 
equal cut-offs for the two ends of the cylinder. 

Vertical engines form an exception to the above statements 
as in these, the cut-off in the crank end of the cylinder should 
occur later than that in the head end by enough to lift the weight 
of the piston, piston rod, crosshead, and connecting rod. 

Another requirement towards securing smooth running is that 
there should be equal leads on the two ends of the cylinder as, 
by this means, the cushioning effect at each end of the piston 
stroke is made equal and, if the lead is sufficient for the size and 
speed of the engine, it will pass the dead centers smoothly and 
without shock. 

It appears from the above discussion that engine valves should 
be set with respect to the cut-off and lead. An inspection of 
Fig. 90 will show that both the cut-off and lead depend upon the 
eccentricity, angle of advance, and steam lap. All of these, 
except the angle of advance, are fixed dimensions and cannot be 
readily changed after the engine is built. However, as valve 
gears are usually constructed, the length of the valve stem may 
be changed or, what amounts to the same thing, the position of 
the valve on the stem may be changed. Changing the length 
16 157 



158 



STEAM ENGINES 



of the valve stem, which moves the valve bodily along on its 
seat, has the effect of increasing the steam lap and decreasing the 
exhaust lap on one end, and decreasing the steam lap and in- 
creasing the exhaust lap on the other end. For example, if the 
length of the valve stem is increased, the steam lap on the head 
end will be increased, the exhaust lap on this end decreased and 
the steam lap on the crank end will be decreased and the exhaust 
lap on this end increased. In setting valves, therefore, the two 
things that may be changed are the angle of advance and the 
length of the valve stem. 




Fig. 94. 



An inspection of Fig. 94 will show that if a valve has equal 
steam laps on head and crank ends, the leads on the two ends will 
also be equal. If the connecting rod was long enough so that 
its angularity did not affect the motion of the piston the cut-off 
on the two ends would also be equal when the steam laps are 
equal but actually, the angularity of the connecting rod does af- 
fect the motion of the piston and causes cut-off to occur later in 
the stroke from head to crank end than it does in the stroke from 
crank to head end. It may be stated as a general rule, therefore, 
that an ordinary slide valve cannot he set to give both equal cut-offs 



VALVE SETTING 159 

and equal leads, although it would be desirable if this could be 
done. 

If the valve gear contains a rocker arm which reverses the 
motion of the valve it is possible to shape the rocker arm so the 
angularity of the connecting rod may be compensated for and the 
cut-offs made equal while retaining equal leads, but this can be 
done for cut-off at only one point of the stroke. If the valve is 
set to cut off at any other part of the stroke either the cut-offs 
or the leads will be unequal. Rocker arms of this kind are not 
made straight, but instead, the part on one side of the pivot is 
made at an angle to the part on the other side of the pivot. 
These are sometimes used on automatic high speed engines but 
are seldom used on plain slide valve engines. 

A plain slide valve may be set to give equal leads on the two 
ends of the cylinder, the cut-offs being unequal, or to give equal 
cut-offs, the leads being unequal, or a compromise may be made 



Fig. 95. 

and the valve set to give slightly unequal leads and slightly 
unequal cut-offs. 

In setting engine valves it is necessary to place the engine on 
dead center. A person cannot judge when an engine is exactly 
on dead center because when near the end of the stroke the crank 
moves through a considerable angle while the piston and cross- 
head move very little. For this reason it is desirable to use some 
method for putting the engine exactly on dead center and the 
following method will be found satisfactory. 

Placing an Engine on Center.^ — An engine is placed on center 
by means of a tram, which is an iron rod pointed at both ends 
and having one end bent at right angles to the main part of the 
rod, as shown in Fig. 95. A tram about 30 inches long and made 
of a rod ^{q inch in diameter is a convenient size to use. 

In placing an engine on center by the use of a tram the fly- 
wheel is turned by hand until the crosshead is within three or four 
inches of the end of its stroke. In turning the engine by hand it 
should always be turned in the direction in which it ordinarily 
runs in order to take up the lost motion or backlash in the various 



160 



STEAM ENGINES 



bearings. When the crosshead has been brought within three 
or four inches of the end of its stroke a scratch mark is made on 
both crosshead and guide, as shown at B Fig. 96, so that by 
bringing the parts of this mark together at any time the crosshead 
will be in the same position as before. The straight end of the 
tram is now placed on a fixed mark on the floor and a mark is 



o 
o 




o 
o 










Fig. 96. 

made with bent end on the rim of the flywheel as shown at A 
in Fig. 96, the crosshead still being in such position that the 
marks on crosshead and guide are together. The flywheel 
is now turned by hand until the crank is past center and the cross- 
head has been brought to the same position as before, as indicated 
by the marks on crosshead , and guide coinciding. The crank 




Fig. 97. 

will now be as far past center as it was in front of center before. 
With the engine in this position the tram is placed on the perma- 
nent mark on the floor and another mark made on the rim of the 
flywheel, as shown at C in Fig. 97. With a tape measure or 
pair of dividers find the point X midway between A and C. Make 
a center punch mark here and turn the engine until the tram, 
still on the permanent mark on the floor, falls square into the 
punch mark, when the engine will be exactly on center. 



VALVE SETTING 161 

The above method should be used in finding the other dead 
center position and, with these two positions marked, the engine 
may be quickly placed on center at any time by turning it until 
the tram, resting on the permanent mark on the floor, comes to 
the punch mark representing the dead center position. 

To Set Valves With Equal Leads. — In setting the valve, two 
results must be accomplished. First, the valve must be made to 
travel equal distances each side of its central position, thus giving 
the valve equal leads on its two sides; second, after making the 
leads equal, they must be adjusted to the desired amount. 

In order to accomplish these results, two adjustments are 
possible; first, the position of the valve on the rod may be 
changed, or, what amounts to the same thing, the length of the 
valve rod may be changed; second, the eccentric may be shifted 
around on the shaft. Changing the length of the valve rod (or 
the position of the valve on the rod) increases the lead at one end 
of the valve and decreases it at the other end. If the rod is length- 
ened the lead will be increased at the crank end and decreased 
at the head end. If it is shortened, the lead will be increased 
at the head end and decreased at the crank end. Shifting the 
eccentric around on the shaft increases both leads or decreases 
both leads, depending upon which direction the eccentric is 
shifted. If the eccentric is shifted so as to decrease the angle 
of advance, both leads are shortened and if the angle of advance is 
increased both leads are lengthened. 

In setting a slide valve, proceed as follows: (1) Set the engine 
on head end dead center and (having removed the cover of the 
valve chest) measure the lead which the valve has on that end. 
If the valve covers the port (negative lead) mark the position 
of the valve and then turn the engine forward until the edge 
of the valve is in line with the edge of the port, and measure the 
distance which the port was overlapped by the valve. 

2. Turn the engine forward until the crank end dead center is 
reached and measure the lead in the same manner. 

3. If the leads are equal and of the required sunount, no further 
adjustment is needed. 

4. If the leads are equal but not of the required amount, move 
the eccentric forward to give more lead or backward to give less 
lead, as required. 

5. If the leads are unequal, they must first be made equal by 
changing the length of the valve rod. To do this, take half of 



162 STEAM ENGINES 

the difference between the leads and change the length of the 
valve rod by this amount, lengthening it if the head end lead is 
larger or shortening it if the crank end lead is larger. This will 
make the leads equal. Then make the leads the required amount 
by the method indicated in (4) above. 

After the valve has been set by measurement, as above, the 
engine should be run and indicator diagrams taken. The indica- 
tor diagrams will show whether or not the valve is set properly, 
and any slight readjustment that may be necessary may be made 
after an inspection of the diagrams. 

The above method of setting valves requires considerable 
turning of the engine by hand which, if the engine is large, may 
be inconvenient. If it is difficult to turn the engine by hand, 
the following method of setting a plain slide valve may be used 
and good results obtained. This method, however, is suitable 
for only plain slide valve engines in which there is no rocker arm 
for equalizing the cut-off on the two ends of the cylinder. 

This method is based on the fact that a slide valve without an 
equalizing rocker arm will give the same maximum port opening 
on the two ends of the cylinder when the leads are equal. The 
valve is therefore first set to give the same maximum port open- 
ing on the two ends of the cylinder. This is done by loosening 
the eccentric on the shaft and turning it around until it gives 
maximum port opening on first one end and then on the other. 
If the maximum port openings are not equal they are made so 
by changing the length of the valve stem by one-half of the 
difference in the maximum port openings. This operation gives 
the valve stem its proper length. The engine is now put on dead 
center and the valve given the proper lead by turning the eccentric 
on the shaft. This adjusts the angle of advance and will 
give equal leads at the two ends of the cylinder. The adjustment 
should now be verified by indicator diagrams as with the preceding 
method. 

Setting Valves for Equal Cut-off. — If there is no equalizing 
rocker arm the steam laps and leads will be unequal when the 
valve is set to give equal cut-off on the two ends of the cylinder 
and, as changing the length of the valve stem is equivalent to 
making the laps unequal, most of the adjustment is made by the 
valve stem. 

The engine is first placed exactly on its head end dead center, 
using a tram for this purpose, as previously described. The 



. VALVE SETTING 163 

eccentric is then loosened and turned on the shaft until the valve 
has the proper lead on the head end. The engine is then moved 
forward until the valve comes line on line with the edge of the 
port, which is its position at cut-off. Now measure the displace- 
ment of the crosshead from the beginning of its stroke. The 
engine is then moved forward again until cut-off occurs on the 
return stroke, and the displacement of the crosshead from the 
crank end of its stroke is measured. If the cut-off on the head 
end is earlier than that on the crank end the valve stem is too 
long, but if the cut-off on the crank end is earlier than that on the 
head end the valve stem is too short. In either case the length of 
the valve stem should be changed by an amount which it is esti- 
mated will correct the inequality in the cut-offs. Changing the 
length of the valve stem will, of course, change the lead on the 
head end; therefore the engine must now be turned to the head 
end dead center and the lead adjusted to its original amount by 
moving the eccentric around on the shaft. The cut-offs are again 
measured for equality and, if necessary, the length of the valve 
stem again adjusted. In setting valves by this method, a valve 
diagram will often prove helpful in determining the amount of 
adjustment to make on the valve stem. 

With all the methods of setting valves described above it is 
necessary to remove the cover of the steam chest. It is con- 
venient sometimes to be able to set the valves without removing 
the steam chest cover and even when steam is on the engine. This 
may be the case with a locomotive when out on the road, due to 
the eccentrics slipping. 

In order to be able to set valves without removing the steam 
chest cover it is first necessary to set the valves properly while 
the steam chest cover is off and then make reference marks on the 
valve stem and steam chest. 

Preparatory to setting the valves by this method a tram with 
both ends pointed and bent at right angles should be made. This 
tram should be about half the length of the valve stem. The 
steam chest cover is then removed and the valve set with equal 
leads by the first method described above. After the valve is 
properly adjusted the engine is placed on the head end dead 
center as shown in Fig. 98, and a punch mark made on the valve 
chest. One end of the tram is placed in this mark and a mark 
made on the valve stem with the other end of the tram. The 
mark on the valve stem is then made permanent by a punch 



164 



STEAM ENGINES 



mark. The engine is then turned to the crank end dead center, 
as shown in Fig. 99, and with the tram in the mark on the valve 
chest, another mark which is also made permanent by a punch 
mark, is made on the valve stem. The valves may now be set 
at any time without removing the steam chest cover by placing 
the engine on center and then turning the eccentric on the shaft 
until the tram reaches from the mark on the valve chest to the 
mark on the valve stem which corresponds to that dead center. 




Fig. 98. 

Types of Slide Valves. — The form of slide valve which has been 
illustrated in the previous discussion is known as the plain or D 
sHde valve, and it has been used in this discussion on account of 
its simplicity. This form of valve is widely used on the cheaper 
grades of slide valve engines; but there are certain objections to 
its use on better grades of engines, the most important of these 
objections being the wire-drawing, or drop in pressure produced 
by it during admission, and its friction. 




Fig. 99. 

Drop in pressure during admission is caused by friction of 
the steam rushing past the edges of the valve and through the 
narrow port opening and by the comparatively slow motion of 
the valve while it is opening and closing. This objection is not 
so serious as might be first thought because while there is a rather 
large drop in pressure, there is also some heat produced by friction 
of the steam which tends to dry the admission steam if it is wet 



VALVE SETTING 165 

or to superheat it if it is dry. However, on the whole, the drop 
in pressure during admission is an objection and some engine 
builders try to avoid it by designing their valves so as to produce 
a large port opening with a small movement of the valve. In this 
way the valve opens quicker and wider and there is less friction 
because of the smaller valve travel. 

One of the ways in which these results have been accomplished 
is by the use of double ported and multiple ported valves. 

The double ported valve illustrated in Fig. 100 is designed to 
give twice as large port opening with the same valve movement as 
would be obtained with the ordinary D valve. Steam surrounds 
the valve and also fills the hollow chambers A and A which extend 




Fig. 100. 

entirely through the valve from one side to the other. The bot- 
toms of the chambers A and A are open so that steam may flow 
from them into the ports. The ports have two openings so 
arranged that when the outside edge of the valve uncovers one 
opening the chamber A or A uncovers the other opening thus 
permitting steam to enter the ports at two points. Exhaust 
occurs past the inside edges of the valve and the inside edges of 
the chambers A and A. Passages are cored over the tops of 
the chambers A and A so that steam exhausted past the in- 
side edge of the valve may find its way to the central exhaust 
chamber. 

The form of valve shown here is partly balanced by means 
of the ring H which slides on the under side of the steam chest 



166 



STEAM ENGINES 



cover. Any leakage of high pressure steam past the ring finds 
its way into the exhaust chamber through the holes M and M. 

This type of valve is widely used on marine engines and it is 
made both balanced and unbalanced. The smaller sizes of valves 
are often unbalanced but the larger sizes are invariably balanced. 

The Trick valve, illustrated in Fig. 101 and named after its 




Fig. 101. 

inventor, is a form of double admission valve, and is used to a 
considerable extent on locomotives. Live steam is admitted 
past the outer edge of this valve and also through the passage A 
cast in the body of the valve. When the outer edge of the valve 
uncovers the port for admission the opposite end of the passage A 
is also uncovered thus giving a double admission of steam, the 




Fig. 102. 

same as with a double ported valve. In a similar way, when the 
outer edge of the valve closes the port for cut-off the opposite 
end of the passage is also closed. The Trick valve gives double 
admission but does not give double exhaust since the passage A 
is not open to exhaust at any time. In this respect the Trick 
valve differs from a double ported valve. 

The Straight-line valve shown in Fig. 102 is both a double 
admission and a double exhaust valve. This result is secured by 
means of two ports through each end of the valve, the port A 



VALVE SETTING 



167 



being for admission, and the port B for exhaust. Both sides of 
the valve are exactly ahke and the balance plate C has recesses 
cored in it to correspond with the port openings in the cylinder 
valve-face. This manner of constructing the pressure plate per- 
mits almost perfect balancing of the valve since both sides of the 
valve are subjected to the same steam pressure. 

The passage A through the valve gives a wide, quick opening 
and thus prevents wire drawing during admission. The passage 




Fig. 103. 

serves the same object with the exhaust, preventing wire drawing 
at release. 

The Straight-line valve is used on automatic high speed 
engines which are controlled by a shaft governor. The shaft 
governor is connected directly to the valve and changes the lead 
and angle of advance. The valve is well adapted for this purpose 
because it is so perfectly balanced that but little power is required 
of the governor in changing its position. 

Multi-ported valves are usually of the gridiron type, consisting 
of a flat plate with a number of slots in it which shdes over a seat 
with a hke number of slots. Valves of this type are illustrated in 
Fig. 103. With this type of valve a large port opening is obtained 



168 



STEAM ENGINES 



with but small valve travel. For example, a valve of this type 
having eight slots and a valve travel of }i in. would have a port 
opening of two inches, while an ordinary valve would have a 
travel of two inches to secure the same amount of port opening. 
The smaller travel of the multi-ported valve reduces its friction 
and the amount of work necessary to move it, and at the same 
time, makes effective lubrication easier. 

Multi-ported valves are usually placed across the cyhnder, as 
shown in Fig. 103, instead of lengthwise of it, and there are usu- 
ally four valves, one for admission to each end of the cylinder and 
one for exhaust from each end. This method of placing the 
valves permits shorter ports and reduces the clearance. 




Fig. 104. 

Some of the sUde valves which have been described and illus- 
trated are balanced or partly balanced. Most balanced valves, 
except piston valves, are balanced by means of a balance plate 
over which the valve shdes, or by means of a balance ring recessed 
into the back of the valve and sUding on the inner surface of the 
steam chest cover plate. 

The excessive unbalanced pressure on the common D-valve 
which causes friction and cutting is mainly due to the large ex- 
haust cavity which is filled with low pressure steam while high 
pressure steam surrounds the outside of the valve. If the valve is 
arranged so that five steam may be admitted to the cylinder from 
the inside of the valve while the outside is subjected to exhaust 
steam the unbalanced pressure is greatly reduced. 

The Ball telescopic valve, illustrated in Fig. 104, is designed on 



VALVE SETTING 



169 



this principle. This valve consists of two parts, one of which 
telescopes into the other. Each part consists of a rectangular 
frame which slides over the ports, and on the back of which is a 
short hollow cylinder. The cylindrical parts telescope and the 
inner one is provided with packipg rings to prevent leakage of the 
steam from the inside to the outside of the valve. 

Steam is admitted to the inside of the valve and the exhaust 
escapes past the outside edges. Both sides of the valve have 
working edges which make unnecessary double ports and large 
clearance volume. 

Piston valves are used on a great many automatic high speed 



^^^^^^ 




^^^^^^^ 



Fig. 105. 

engines. This type of engine requires a balanced valve because 
the governor is attached directly to the valve and governs the 
speed by changing the position of the valve. If a large amount of 
power is required to move the valve a sensitive governor cannot 
be used with it. 

Piston valves are cylindrical valves moving in the direction of 
their axis. They may be made either for inside or for outside 
admission. The steam ports consist of annular spaces surround- 
ing the valve, and the admission and exhaust edges extend all 
around the circumference of the valve, hence a large port opening 
is secured with a small diameter of valve. 

Packing rings to prevent leakage of steam are used on the 
larger sizes of piston valves but the smaller sizes have none. The 



170 STEAM ENGINES 

Ideal piston valve illustrated in Fig. 105 shows one method of 
using packing rings on piston valves. Only one end of the valve 
is shown here, the other end being exactly like this one. Two 
rings A are used on each end of the valve and they are made the 
exact size of the cylinder bore. These are not ''spring" rings, as 
they would cut the cyhnder, but they are spht and are slightly 
adjustable by means of the four shoes C. Thus the rings may be 
adjusted to take up wear and keep the valve steam tight at all 
times. 



CHAPTER XIII 
SHIFTING ECCENTRIC AND MEYER VALVE 

Shifting Eccentric. — In the plain slide valve engine the eccen- 
tric is fastened to the shaft by means of a set screw or key, hence, 
the cut-off occurs at a fixed point in the stroke and cannot be 
changed unless the engine is stopped and the eccentric moved 
around on the shaft. This would change the angle of advance 
and consequently the point of cut-off. 

An inspection of the valve diagram, Fig. 88, will show that 
the point of cut-off with a plain slide valve may be changed by 
changing either the angle of advance or the eccentricity. Cut- 
off may be made earlier either by increasing the angle of advance 
or decreasing the eccentricity and it may be made later either 
by decreasing the angle of advance or increasing the eccentricity. 

An eccentric constructed in such manner that its eccentricity 
and angle of advance may be changed without stopping the engine 
is used on the automatic high speed type of engine. This type of 
eccentric, which is called a shifting eccentric, is attached to the 
governor in such manner that the position of the governor controls 
the eccentricity and angle of advance and by this means regulates 
the speed of the engine by changing the point of cut-off to suit the 
load carried by the engine. A device of this kind makes it 
possible to use the simple slide valve and, at the same time, secure 
a variable cut-off in governing the speed of the engine. 

The principle of the shifting eccentric is illustrated in Fig. 
106. The eccentric is not fastened directly to the shaft but to 
the side of a plate, C, with a projecting arm. The projecting 
arm is pivoted at A to a point on one of the spokes of the flywheel. 
The governing mechanism is contained in the flywheel and is 
attached to the eccentric on the side opposite the projecting arm, 
C, hence, the eccentric turns with the flywheel. The shaft 
passes through a slot cut in the eccentric, the width of the slot 
being a little greater than the diameter of the shaft. The slot 
is curved to a radius OA equal to the distance from the pivot A 

171 



172 



STEAM ENGINES 



to the center of the shaft, 0, so that the eccentric is free to swing 
about the pivot, A, without touching the shaft. 

The governor, which is attached to the eccentric, changes 
the position of the eccentric with respect to the shaft and thus 
changes the eccentricity and angle of advance. For example, 
when the eccentric and shaft occupy the positions shown by the 
full lines in Fig. 106 the eccentricity, which is the distance from 
the center of the shaft to the center of the eccentric, is OE and the 
angle of advance is the angle BOE. If the eccentric is shifted 




Fig. 106. 

until the shaft occupies the position indicated by the dotted 
circle, with its center at 0', the eccentricity will be O'E and the 
angle' of advance will be the angle BO'E. It will be observed 
that in shifting from the first to the second position the eccen- 
tricity has been decreased and the angle of advance increased. 
Both of these changes have the effect of making the point of 
cut-off occur earher in the stroke, as shown by the valve diagram 
in Fig. 107. In this valve diagram the full lines represent the 
first position of the eccentric, with cut-off occurring at C, and 
the dotted, hues represent the second position of the eccentric 



SHIFTING ECCENTRIC AND MEYER VALVE 173 

with the shorter eccentricity and greater angle of advance, and 
the cut-off occurring at C. 

The manner in which a swinging eccentric on an automatic 
high speed engine is operated by the governor is illustrated in 
Fig. 108. The eccentric with the slot and shaft, a, is shown at R. 
This is fastened to a plate T which is pivoted to the flywheel at 
S so that it may turn about the point S and thus change the 
eccentricity and angle of advance. The governor consists of the 
weight C, the link H, and the spring E. As the flywheel turns, 




Fig. 107. 

the weight, C, is acted upon by centrifugal force which causes 
it to move outward from the center of the flywheel. The out- 
ward movement of the weight is resisted by the spring E so that 
for any particular speed the weight C will move outward until 
the resisting force of the spring just balances the centrifugal 
force acting upon the weight. The weight, C, is pivoted to the 
arm of the flywheel at 0, and the link H is pivoted to the eccen- 
tric at T. When the weight, C, moves outward the pivot T is 
moved in an opposite direction and the center of the eccentric 
is moved nearer to the center of the shaft. This decreases the 
eccentricity and increases the angle of advance^ which causes 

17 



174 



STEAM ENGINES 



cut-off to occur earlier in the stroke and reduces the volume of 
steam supplied to the C3^1inder of the engine. A movement of the 
weight, C, towards the center moves the eccentric in the opposite 
direction, lengthening the cut-off and admitting a greater volume 
of steam to the cylinder. 

The amount of centrifugal force acting on the weight, C, 
increases with an increase in speed and decreases with a decrease 
in speed of the engine. If, on account of an increase in load on 
the engine, the speed should be decreased the centrifugal force 
acting on C would become smaller and the spring would pull C 
towards the center of the shaft and move the center of the eccen- 




FiG. 108. 

trie, n, away from the center of the shaft, a; thus cut-off would 
be made later and an increased amount of steam would be 
admitted to the engine to make it go faster. So, also, if on 
account of a decrease in the load the speed of the engine should in- 
crease, the centrifugal force would become greater, and C would 
move farther away from the center of the shaft thus moving the 
center of the eccentric towards the center of the shaft, making 
cut-off earlier and reducing the amount of steam admitted to the 
engine. 

The spring, E, acts as a controlling force upon the weight, C, 
and regulates its position for any given speed, therefore the stiff- 
ness of the spring controls the speed at which the engine will run. 



SHIFTING ECCENTRIC AND MEYER VALVE 175 

The stiffness of the spring may be changed by adjusting the 
length of the connection, P, between the spring and the weight 
C, which is provided with a turn buckle for that purpose. 

The effects of different positions of the swinging eccentric 
upon the distribution of steam to the cyhnder may be studied 
from a valve diagram such as that shown in Fig. 109. The swing- 
ing eccentric gives its greatest eccentricity when the engine is at 
rest or when the load is the greatest. In the valve diagram, 
Fig. 109, the line OE represents this eccentricity and the valve 
circle shows the cut-off occurring at ^ i. The arc ED is drawn with 
a radius equal to the length of the arm on which the eccentric 




Fig. 109. 



swings, or AO in Fig. 106. As the cut-off is shortened by the gov- 
ernor, the eccentricity becomes smaller and for any position of the 
governor the eccentricity will be the distance from the center 
of Fig. 109 to the arc ED. Thus, when the eccentric has been 
moved one-half of its total swing, its eccentricity will be OE', 
the point E' being located halfway between E and D, and the 
corresponding cut-off will occur when the crank is in the position 
OAi' A valve circle drawn on OE' will give, by its intersection 
with* the lap circle, the point of cut-off A 2 for the new position 
of the eccentric, since the lap circle is the same for all positions 
of the eccentric. In a similar manner, the cut-off for any position 
of the eccentric may be found by drawing a line from to a 



176 STEAM ENGINES 

point on the arc ED which represents the position of the eccentric, 
and then drawing the valve circle on this line. 

The events occurring past the exhaust side of the valve may- 
be located by extending the valve circle diameters through the 
center and drawing valve circles on them of the same size as 
those used in locating the point of cut-off. In this way, Fig. 
109 show^s that the point of compression occurs at Ki when the 
cut-off occurs at Ai and that the point of compression occurs at 
K2 when the cut-off occurs at A 2. 

It will be observed from the valve diagram that the lead in- 
creases as the cut-off is shortened or the load becomes lighter, 
and that the amount by which the lead changes depends upon 
the length of the projecting arm to which the eccentric is fastened. 
When the cut-off occurs at A 1 the amount of lead is LM and when 
the cut-off occurs at A 2 the lead is LN. The automatic high 
speed engine is designed to run at practically constant speed and 
to change the point of cut-off to suit the load carried. The lead 
has a cushioning effect upon the piston, hence, at constant speed 
the lead should be practically constant or if it changes at all it 
should increase when the point of cut-off occurs later in the stroke, 
as the engine is then carrying its heaviest load. The fact that 
the swinging eccentric increases the lead for short cut-offs is 
sometimes urged as an objection to this method of governing. 
The lead may be made more nearly constant by pivoting the 
eccentric further from the center of the flywheel since this gives 
a flatter arc ED on the valve diagram, but there will always be 
some change in lead with this kind of eccentric. 

A study of the valve diagram- shows also that the period of 
compression is increased as the cut-off becomes shorter. Com- 
pression cushions the piston, hence it would be desirable to have 
a constant amount of compression on a constant speed engine. 
However, the changes in compression have but little effect upon 
the smoothness of running of the automatic high speed engine 
on account of the large clearance volume which the engines have. 
The large clearance volume, made necessary by the high speed 
of the engine, flattens the compression curve on the indicator 
diagram, showing that the cushioning effect upon the piston is 
applied gradually, hence a change in the point of compression 
does not affect the cushioning effect as much as it would on an 
engine with smaller clearance. 

The action of the flywheel governor and swinging eccentric 



SHIFTING ECCENTRIC AND MEYER VALVE 177 

may be seen from Fig. 110 which shows a series of indicator 
diagrams drawn for different cut-offs. These diagrams were 
drawn when the engine was running at the same speed, but with 
different loads. The changes in lead are particularly noticeable 
on these diagrams, the one with the latest cut-off showing almost 
no lead and the others showing increasing lead as the cut-off 
is shortened. The manner in which the point of compression 
is changed with the cut-off is also very noticeable. With the 
latest cut-off the point of compression occurs almost at the end 
of the exhaust stroke, but, with no load, when cut-off is earliest, 
the point of compression occurs before mid-stroke. 

The valves used on automatic high speed engines are always 
balanced valves, balancing being secured either by a cover plate 
on the back of the valve or by the use of a cylindrical or piston 
valve. It is very necessary, in these engines, that the power 




Fig. no. 

required to move the valve back and forth be reduced to a mini- 
mum because the valve is moved by the governor and, if much 
power is required to move it, the sensitiveness of the governor 
will be reduced. For this reason the valves on automatic high 
speed engines are balanced and friction is reduced as much as 
possible by good workmanship in making the valves. 

An automatic high speed engine is built to run at a certain 
speed and the governor is designed and adjusted by the manufac- 
turer for this speed. The speed may, however, be changed to a 
certain extent either by changing the tension of the springs or by 
decreasing the weights. The speed will be increased by increas- 
ing the tension of the springs or by decreasing the weights as, 
in either case, a greater speed will be required to overcome the 
tension of the springs. 

Effects Produced by Slide Valve. — One of the principal objec- 
tions to the plain slide valve engine is its unfavorable steam distri- 



178 



STEAM ENGINES 



bution. All of the events, admission, cut-off, release, and com- 
pression are controlled by a single valve and if one of the events 
is changed all of the others are changed in proportion. For 
example, if cut-off is made to occur earlier in the stroke all of the 
other events, release, compression, and admission are also made 
to occur earlier. 

The construction of a plain slide valve is such that the angle 
turned through by the flywheel during expansion is always equal 
to the angle turned through during the period of compression, as 
an inspection of the valve diagram for this type of valve will 
show. Ordinarily a long period of expansion is desirable as a 
greater expansion of the steam is then secured and more work 
obtained from the steam, but this can be obtained, with the plain 



\ 



Fig. 111. 



slide valve, only by having a long period of compression, which 
reduces the amount of work secured from the steam. With these 
opposing conditions the best results will be obtained from the 
engine by locating the point of cut-off early enough so there will 
be some expansion of the steam but not enough to result in 
excessive compression. For this reason the point of cut-off 
for a plain slide valve engine is usually made to occur between 
one-half and seven-eighths of the stroke. If it occurs earlier than 
mid-stroke the compression will be excessive, and if it occurs 
later than at seven-eighths of the stroke there will be but httle 
expansion of the steam. Under the latter condition the pressure 
of the steam at release will be almost the original pressure during 
admission and when the exhaust valve is opened this pressure 
will be wasted. 



SHIFTING ECCENTRIC AND MEYER VALVE 179 

Meyer Valve. — A study of the plain slide valve shows that if 
the point of cut-off could be changed without changing the other 
events, a much more economical steam distribution could be 
obtained by cutting off the steam earlier in the stroke and 
allowing it to expand a greater number of times. This is made 
plain by considering Fig. Ill in which the full line indicator 
diagram is from a plain slide valve engine and the dotted Hne 
shows the shape of diagram that would be obtained if the point 
of cut-off was made earlier without disturbing any of the other 
events. These diagrams show that more than half as much work 
is done with the early cut-off as with the late one, but that less 



- ThLn* 




Fig. 112. 



than half as much steam is admitted to the cylinder. Therefore' 
more work is obtained per pound of steam with the early cut-off 
than with the late one. This result is to be expected since very 
little of the expansion force of the steam is used when a late cut- 
off is employed. 

An early cut-off by a slide valve may be obtained by means 
of a device called a Meyer valve. The Meyer valve consists of an 
auxihary valve sliding on the back of the regular or main shde 
valve, as illustrated in Fig. 112. The main valve is like a plain 
slide valve except that instead of admitting steam past the outer 
edge, the valve is extended and has two ports, P and Pi through it 
and the admission steam flows through these ports. The valve 



180 STEAM ENGINES 

is constructed in this way in order to provide a proper surface 
on which the auxiliary or riding valve may slide. 

The auxiliary valve consists of two blocks, Vi and V2, carried 
on the valve rod K which is operated by a separate eccentric. 
The auxiliary valve slides on the back .of the main valve and, at 
the proper instant, covers the ports, P and Pi through the main 
valve thus stopping the flow of steam through them and cutting 
off the steam from the cylinder. The auxiliary valve will close 
the ports P and Pi through the main valve and cause cut-off 
when it has moved with respect to the main valve a distance 
equal to e, in Fig. 112, which is called the clearance. The clear- 
ance may be adjusted by turning the handwheel, W, which turns 
the valve rod. The valve rod is provided with left- and right- 
hand threads so that turning it moves the blocks Vi and V2 
farther apart or closer together and thus changes the clearance e. 

The main valve is set the same as an ordinary slide valve but 
with a late cut-off in order to secure a short compression. The 
eccentric which operates the auxiliary valve is then set, usually at 
180° from the crank, and the blocks placed on the valve stem in 
such position as to give equal cut-offs on the two ends. This is 
done by turning the blocks around separately on the valve stem 
until the clearance, e, is adjusted properly to equalize the cut-offs. 
The point of cut-off is then adjusted for both ends of the cylinder 
by turning the valve stem. 

The auxiliary valve controls only the cut-off and it does this 
by closing the ports through the main valve before the main 
valve itself cuts off in the regular way. The main valve controls 
all the other events, release, compression, and admission. By 
setting the main valve to give a short compression and setting 
the Meyer valve to give an early cut-off, an indicator diagram 
may be obtained which resembles very closely the indicator 
diagram obtained from a Corliss engine, as shown by \he dotted 
diagram in Fig. 111. 

The action of the Meyer valve may best be examined by means 
of a valve diagram such as that shown in Fig. 113 In this 
diagram the cricle A is for the main valve and the circle B is for 
the auxiliary valve. The diameter of A is located so that the 
angle JOF is equal to the angle between the crank and main 
eccentric but laid out in a direction opposite to the rotation 
as indicated by the arrow. Likewise the diameter of the auxiliary 
valve circle, OG, is laid out so that the angle JOG is equal to the 



SHIFTING ECCENTRIC AND MEYER VALVE 181 

whole angle between the crank and auxiliary eccentric but in a 
direction opposite to the rotation. It will be observed that while 
the auxiliary eccentric is usually placed directly opposite the 
crank, in this case it is placed at a somewhat greater angle than 
180° in order to make the diagram more general. Both the main 
and auxiliary eccentrics in this case have the same eccentricity 
as they usually have in practice but this is not at all necessary. 

Since both valves are moved by eccentrics, their relative 
motion with respect to each other may be represented by a third 
valve circle C which is drawn with its diameter OK equal and 
parallel to GF. The position of the crank at cut-off for any 
amount of clearance may then be located on the circle C by 




Fig. 113. 

drawing a line OE through so that the distance OD is equal to 
the clearance. Cut-off will then occur when the crank is in the 
position OE. 

The circle C can be used only for finding the point of cut-off 
on one end of the cylinder. In order to find the cut-off on the 
other end the line KO must be extended and another circle of 
the same size as C drawn on it. The cut-off for the other end of 
the cylinder may then be located on it in the same manner as just 
described. Since admission, release, and compression are con- 
trolled entirely by the main valve they must be located by means 
of the main valve circle. Admission is located by drawing the 
lap circle in A as with other valve diagrams that have been de- 
scribed before. Compression and release must be located by 
extending OF and drawing on it another valve circle of the same 
size as A. By drawing the exhaust lap arc in this circle, both 

18 



182 STEAM ENGINES 

release and ompression may be located the same as on other 
valve diagrams. 

The clearance of the Meyer valve may be adjusted to cut off 
the supply of steam at any point between the beginning of the 
stroke and the point at which the main valve cuts off, the only 
limitations being that the valve must give sufficient port opening 
at the beginning of the stroke and must not re-open before the 
main valve closes, a matter which depends upon the design of the 
valve. 

Engines which have Meyer valves are usually governed by 
means of a throttling governor which reduces the pressure in the 
steam chest. Attempts have been made to operate the auxiliary 
valve from a shifting eccentric connected to a flywheel governor 
as is done in the automatic high speed engine but this method is 
not successful with the Meyer valve because it gives a very 
unfavorable steam distribution. If the swinging eccentric 
lengthens the cut-off on one end of the cylinder it shortens it on 
the other because such a governing device moves the auxiliary 
valve as a whole, whereas, in order to change the cut-off equally 
on both ends of the cylinder, the blocks must be separated or 
brought closer together. This can only be done by turning the 
valve rod, and a swinging eccentric cannot do this. 

The Meyer valve finds its most successful use on engines 
which carry a fairly uniform load such as on those running air 
compressors and similar machines. For this class of service the 
point of cut-off may be adjusted by hand to the most favorable 
part of the stroke, which will be as early as the load will permit, 
and then the small fluctuations in the load may be taken care of 
by the throttling governor, and a constant speed maintained. 
The point of cut-off may be changed while the engine is running, 
since the handwheel is outside the valve chest, hence, if the load 
changes enough to require a different point of cut-off, it may be 
changed to suit the new conditions. The handwheel is usually 
provided with a pointer and scale marked in the fractions of the 
stroke at which cut-off occurs, so that the clearance may be 
adjusted to any point of cut-off by simply turning the handwheel 
until the pointer is at the fraction at which it is desired that cut- 
off shall occur. 



CHAPTER XIV 
REVERSING MECHANISMS 

Reversing Gears. — Many kinds of steam engines require a 
form of valve gear by means of which the direction of rotation 
may be readily reversed. Some of the kinds of engines which 
require a reversing valve gear are locomotives, marine engines, 
hoisting and winding engines, traction engines, and rolling mill 
engines. 

A study of the valve diagram shows that any slide valve 



CRANK 




Fig. 114. 

engine may be reversed by simply shifting the eccentric around on 
the shaft. For an outside admission valve and no reversing 
rocker arm between the eccentric and the valve, the eccentric 
is set ahead of the crank by an angle somewhat greater than 90°. 
If such an engine is to run in a clockwise direction, the relative 

CRANK 




Fig. 115. 

positions of the crank and eccentric are as shown in Fig. 114. 
If it is desired to have this engine run in a counterclockwise 
direction, the eccentric must be loosened and turned to the 
position shown in Fig. 115, when the eccentric will be ahead of 
the crank but in a counterclockwise direction. This method 
of reversing an engine could not be used with any engine which 
19 183 



184 STEAM ENGINES 

required a quick reversal because of the time required to change 
the position of the eccentric and also because this method makes 
it necessary first to stop the engine. 

Many forms of valve gears have been invented which will 
quickly reverse the engine without having to change the position 
of the eccentric. Some of these mechanisms make use of two 
eccentrics, one for each direction of rotation; some of them use 
only one eccentric; and some derive their motion from a pin on 
the end of the shaft placed with its center a short distance from 
the center of the shaft so as to give the same motion as a short 
crank. This arrangement is equivalent to a single eccentric. 
Some of the more common forms of reversing mechanism are the 
Stephenson link motion, the Walschaert valve gear, and the 
Woolf reversing gear. 



CRANK 




Fig. 116. 

Stephenson Link Motion. — The Stephenson link motion is a 
reversing mechanism made up of two eccentrics, two eccentric 
rods connected by a link and a single valve stem. The two eccen- 
trics are keyed or fastened to the shaft in the positions shown in 
Fig. 116. The eccentric OE is in the proper position for producing 
clockwise rotation and the eccentric OE' is in the proper position 
for producing rotation in a counterclockwise direction. The 
mechanism is arranged so that either of these eccentrics may be 
made to operate the valve, and rotation in either direction pro- 
duced at will. 

A Stephenson link reversing gear is illustrated in Fig. 117 
showing the relations of the different parts. The two eccentrics 
are shown at A and B and the crank at C. It will be observed 
that the positions of the two eccentrics in relation to the crank 



REVERSING MECHANISMS 



185 




2 



186 STEAM ENGINES 

is the same as those shown in Fig. 116. The eccentric A is in the 
proper position for producing rotation in a clockwise direction, 
and B is in the proper position for producing rotation in a counter- 
clockwise direction. Each eccentric has its own eccentric rod 
connected to one end of the slotted link L. The link L is sus- 
pended by M from the bell crank A^ which is pivoted at R, 
The bell crank is operated by the lever P so that by moving P 
the link L may be raised or lowered. The end of the valve stem 
V is bolted to a block W which fits in the slot of the link but is 
free to move in it, so that, as the link is raised or lowered, the 
block W and valve stem remain stationary. In this way the 
valve stem may be brought into line with either eccentric rod or 
it may be given any intermediate position between them. 

When the block is at the end of the link, and the valve stem 
in line with the eccentric rod E, as shown in Fig. 117, the valve 
has the same motion as if connected directly to the eccentric 
A and the engine will rotate in a clockwise direction. In this 
case the eccentric B has no effect upon the motion of the valve. 
Its only effect is to swing the bottom of the link about the point 
a as a pivot in the same manner that a pendulum swings. 

When the link is raised so that the valve rod is in line with 
the eccentric rod F, all of the motion of the valve comes from 
the eccentric B and the engine therefore runs in a counter- 
clockwise direction. In this case the only effect of the eccentric 
A is to swing the upper end of the link like a pendulum about the 
block TF as a pivot (the block now being at the lower end of the 
link). 

When the link is raised so that the block is mid-way between 
the two ends of the link, the valve is acted upon equally by both 
eccentrics, one tending to produce clockwise rotation and the 
other tending to produce counterclockwise rotation; therefore, 
the engine will not run in either direction. 

The valve mechanism is said to be in '^mid-gear" when the 
block is in the middle of the link and to be in ''full gear" when 
the block is at the end of the link. There are two ''full gear" 
positions, one called "full gear foward" and the other "full gear 
backward," to indicate the position of the link which will cause 
the engine to move forward or backward. 

The effects upon the valve motion of placing the block in 
different positions in the link and also the effects upon the steam 
distribution to the cylinder may be studied by means of a valve 



REVERSING MECHANISMS 



187 



diagram. A valve diagram for a Stephenson link motion is 
shown in Fig. 118. In drawing this diagram it must be remem- 
bered that when the valve mechanism is in "full gear/' the motion 
of the valve is the same as if it were connected to one of the 
eccentrics directly and the other eccentric were not present; 
therefore the line OE represents the eccentricity of one of the 
eccentrics, and the angle COE represents the angle of advance 
of this eccentric. The valve circle OAE is then drawn on the 
line OE and the steam lap circle AHJB is drawn about as a 
center. The line AF then gives the position of the crank at 




cut-off when the valve mechanism is in full gear and the engine 
is to rotate in a clockwise direction. 

Both eccentrics of a Stephenson link motion have the same 
eccentricity and the same angle of advance, hence the location of 
the crank for full gear cut-off when the engine is rotating in a 
counterclockwise direction is found on the valve circle OHE'B. 
The diameter of this valve circle is the same as OE and the angle 
of advance DOE' is the same as the angle COE. Since the valve 
has the same steam lap for any position of the link, cut-off will 
occur when the crank is in the position OG. 



188 STEAM ENGINES 

When the block is at any position in the link intermediate 
between its two ends, the motion of the valve is derived from 
both eccentrics and this motion might be produced by a single 
imaginary eccentric having a smaller eccentricity and a greater 
angle of advance than either of the actual eccentrics. 

Op. the valve diagram the center of the eccentric is at E for. 
one full gear position and at E' for the other full gear position. 
As the link is shifted, the center of the imaginary equivalent 
eccentric moves along a curved path EPE' which is approximately 
the arc of a circle. The arc EPE' may be drawn as follows: 
Take a radius equal to 

in which s is the distance from one eccentric center to the other 
(the distance from E to E' on the valve diagram), I is the length 
of the eccentric rods, and k is the length across the link measured 
from the center of the block in one full gear position to the center 
of the block in the other full gear position. With a center on the 
line KL and a radius as calculated above draw an arc passing 
through the points E and E', cutting KL at the point P. A 
valve circle drawn with OP as a diameter gives the location of 
the crank at cut-off OM for the mid-gear position of the valve 
mechanism. 

For any position of the link intermediate between full gear 
and mid-gear the point of cut-off may be located by drawing a 
valve circle for that position of the link. This is done by dividing 
the arc EPE' in the same proportion that the link is divided by 
the position of the block. For a position of the block half way 
between the mid-gear and the full gear positions, the valve 
circle is drawn on OR as a diameter, R being half way between 
E and P, and it is seen that cut-off occurs when the crank is in 
the position ON. 

It will be observed from Fig. 118 that the latest cut-off is 
obtained with the link in its full gear position and that as the 
link is brought towards the mid-gear position the cut-off becomes 
earlier, being at ON for one-quarter gear, and at OM for mid- 
gear. The Stephenson link motion may therefore be used to a 
certain extent as a governor since the point of cut-off may be 
regulated by it. On locomotives, where this form of valve gear 
is commonly used, the speed is regulated by both the throttle 
valve and the valve gear. In starting a locomotive pulling a 



REVERSING MECHANISMS 189 

load the link motion is put in full gear where the latest cut-off 
is obtained. The full steam pressure then acts upon the piston 
for nearly the entire stroke. As the locomotive comes up to 
speed, the link motion is gradually notched up towards mid-gear 
which shortens the cut-off and allows more of the expansive force 
of the steam to be used. 

The valve diagram shown in Fig. 118 is not complete, only 
enough of it being shown to illustrate the method of drawing it. 
The remainder of the diagram is omitted in order not to complicate 
the figure. For the full gear position the line OE would be 
extended to the other side of the center E and another valve 
circle drawn on it with the exhaust lap to locate the positions 
of the crank at release and compression, in the same manner as 
for an ordinary slide valve diagram. These two valve circles 
will give all of the full gear events for one end of the cylinder. 
The events for the other end of the cylinder would be found by 
drawing an exhaust lap circle in the valve circle EO and a steam 
lap circle in the other valve circle. The same process would be 
followed in drawing a complete diagram for any other position 
of the link. 

The link motion illustrated in Fig. 117 is known as the open 
rod construction and the valve diagram shown in Fig. 118 is 
for this form of link motion. There is another form of link 
motion called the crossed rod construction in which the eccentric 
rod E, Fig. 1 17, is attached to the end/ of the link, and the eccen- 
tric rod F is attached to the end d of the link. With either the 
open or crossed rod construct on the eccentric rods are alter- 
nately open and crossed every half revolution, but these two 
constructions may be distinguished by means of the following 
rule: For an outside admission valve and no reversing rocker 
arm the link motion is open rod construction if the rods are open 
when the crank is in the dead center position on the side of the 
shaft away from the valve. Under the same conditions the 
valve gear is of the crossed rod construction if the rods are 
crossed when the crank is in dead center position on the side 
away from the valve gear. 

The motion of the valve for crossed rods is very different from 
its motion with open rods. For this reason, in dismounting a 
Stephenson link motion it is very necessary to know whether the 
link motion is open or crossed rod construction and, in reassem- 
bling, not to change from one construction to the other. A center- 



190 



STEAM ENGINES 



line diagram of a crossed rod link motion is illustrated in Fig. 
119. 

The valve diagram for the crossed rod construction is the 
same as that for the open rod construction except that the arc 




Fig. 119. 



ERPE' Fig. 118 is drawn so as to curve in the opposite direction. 
This is done by placing the center of the arc on the opposite side 
of the center of the diagram. The method of finding the length 
of radius used for drawing the arc with both the open and crossed 




rod constructions is the same. The valve diagram for crossed 
rods is illustrated by Fig. 120. 

It will be observed from Figs. 118 and 120 that with the open 
rod construction the lead increases as the link is moved from 



REVERSING MECHANISMS 



191 



full gear to mid-gear while with the crossed rod construction the 
lead decreases as the link is moved from full gear to mid-gear. 
The fact that the lead changes for different positions of the link 
is sometimes urged as a disadvantage of the Stephenson link 
motion, and some of the later types of reversing mechanism are 
designed to give constant lead. For use on locomotives, if the 
lead must change, it is desirable to have it increase from full gear 
to mid-gear in order to give more cushioning effect as the engine 
speeds up. 

In setting the valve of a Stephenson link motion the link is 
placed in full gear and the same method is then used as in setting 
the ordinary slide valve. The precaution must be taken however 
to see that the eccentric rods are of the same length; otherwise 
the steam distribution will not be the same for forward running 
as for backward running. 

Walschaert Valve Gear. — While the Stephenson link motion 
is used on most American locomotives at the present time, it is 




Fig. 121. 



being displaced on the later types of locomotives by a form of 
reversing mechanism called the Walschaert valve gear. 

The Stephenson link motion and other types of link motions 
have two eccentrics, one for forward motion and the other for 
backward motion. The Walschaert reversing mechanism, and 
others of the same type have but one eccentric or, in some cases, 
only a pin on the end of the crank shaft which acts as a crank 
of small throw. Reversing mechanisms having but one eccentric 
are called radial valve gears. 

With the Walschaert valve gear, the valve takes part of its 
motion from the crank shaft and part from the crosshead, these 
two motions being combined into the actual motion of the valve. 
The part of the motion that comes from the crank shaft is pro- 
duced by connecting the valve to a pin set at 90° to the crank. 
A valve connected directly to a pin placed at 90° to the main 
crank is shown in Fig. 121 in which OC is the crank and E is the 



192 



STEAM ENGINES 



pin which operates the valve. The small crank OE may be 
spoken of as an eccentric since it serves all the purposes of an 
eccentric. 

An engine fitted with a mechanism like that shown in Fig. 
121 will not run if the valve has steam lap, because, when the 
piston is at the end of its stroke the valve is in its mid-position 
and covers the steam ports a distance equal to the steam lap. 
In order to allow the engine to run the valve must be displaced 
a distance equal to the steam lap when the crank is in the position 
shown in Fig. 121. The valve will then be on the point of open- 
ing when the piston reaches the end of its stroke. For proper 
operation the valve should also be given some lead; that is, 
when the piston is at the end of its stroke, as shown in Fig. 121, 
the valve should be displaced from its mid-position a distance 
equal to the steam lap plus the lead. 




Fig. 122. 



The method of securing lead with the Walschaert valve gear 
is shown in Fig. 122. The valve stem is not connected directly 
to the eccentric rod but to a combining lever which is connected 
to the crosshead. The eccentric rod is also attached to the com- 
bining lever, so that its swing is due both to the motion of the 
crosshead and to the eccentric crank OE. The crosshead carries 
a drop bar BD fastened rigidly to it and the combining lever is 
connected to it by a union link. The use of the union link is 
made necessary by the fact that the crosshead travels in a 
straight line while the lower end of the combining lever swings 
in the arc of a circle. The rise and fall of the lower end of the 
combining lever is therefore taken up by a slight swing of the 
union link about the point D, thus permitting the valve stem to 
travel in a straight line through its stuffing box. 

The combining lever is supported at the point where the 
eccentric rod is attached to it and this point forms the fulcrum 



REVERSING MECHANISMS 



193 



of the lever. The motion of the crosshead then displaces the 
valve independently of the motion produced by the eccentric 
crank. The combining lever is in a vertical position when the 
crosshead is at the middle point of its travel; therefore, when the 
piston is at either end of its travel the valve is displaced from its 
mid- position a distance depending upon the distance from the end 
of the combining lever to the point at which the eccentric rod is 
attached. This distance is made great enough to displace the 
valve a distance equal to the steam lap plus the lead, as shown in 
Fig. 122. When the crank is on the back dead center, as shown 
in Fig. 122, the combining lever displaces the valve to the right 
of its mid-position and when the crank is on the other dead 




Fig. 123. 

center, the valve is displaced to the left of its mid-position, thus 
giving lead at both ends of the piston stroke. 

By erecting the mechanism so that the combining lever stands 
in a vertical position when the piston is at the middle point of its 
stroke, the displacement of the valve is made equal for each dead 
center position of the crank. If the valve has the same steam 
lap on both sides, the leads will then be equal. Moreover the 
lead will be constant since it depends only upon the proportions 
of the combining lever, which are fixed. 

While the valve gear illustrated in Fig. 122 gives the valve 
the proper lead and will run the engine correctly, the direction of 
rotation of the engine cannot be reversed nor can the point of 
cut-off be changed. Both of these objects are accomplished in 
the Walschaert valve gear in the manner illustrated in Fig. 123. 
Instead of a single eccentric rod extending from the eccentric 



194 



STEAM ENGINES 



crank to the combining lever, the eccentric rod is divided into 

two parts and a Hnk 



interposed between 
them. The link is 
supported at its center, 
about which it oscillates. 
The part of the eccentric 
rod which extends from 
the link to the combin- 
ing lever (called a radius 
bar) has a block on its 
end which may be moved 
up or down in the link. 
Moving the block from 
one end to the other of 
the link reverses the 
movement of the valve 
and therefore reverses 
the direction of rotation 
of the engine. The 
point of cut-off is 
changed as the block is 
moved from the center 
of the link towards the 
end. The latest cut-off 
occurs when the block is 
at the end of the link for 
then the valve travel is 
greatest and the valve 
uncovers the steam 
ports the greatest 
amount. As the block 
is moved towards the 
center of the link, the 
valve uncovers the ports 
a smaller amount and 
the cut-off is shortened. 
By curving the link 
to a radius equal to 
the length of the radius 
bar the lead is kept con- 




REVERSING MECHANISMS 195 

stant for all positions of the block in the link. This is one of 
the points of difference between the Stephenson link motion and 
the Walschaert valve gear. The former varies the lead for 
different points of cut-off and the latter keeps the lead constant. 
A valve gear may be tested for constant lead by shifting from 
one full gear position to the opposite and watching the valve or 
valve rod. With constant lead the valve will not move as the 
mechanism is shifted from one full gear position to the other. 

An example of the Walschaert valve gear as applied to a 
locomotive is illustrated in Fig. 124. It will be observed that the 
eccentric rod is not connected directly to the lower end of the 
link but rather to an arm which projects from the lower end of the 
link and which inclines backward. This is made necessary by 
the fact that the eccentric rod is not horizontal but inclines at a 
considerable angle. This angularity of the eccentric rod distorts 
the motion of the valve. The projecting arm on the link is added 
to bring the eccentric rod as near the horizontal as possible; and 
the arm is shaped so as to correct the distortion in the valve 
motion produced by the angularity of the eccentric rod. 

The Walschaert valve gear is not a new device since, it was 
invented in the year 1848 and has been in more or less extensive 
use on European locomotives since then. It is only within recent 
years, however, that it has been used to any extent in the United 
States. Its many advantages over the Stephenson link motion is 
now causing its rapid adoption on locomotives in this country. 

The Walschaert gear is much lighter in weight than the 
Stephenson link motion. The Stephenson link motion has two 
heavy eccentrics with their straps and eccentric rods while the 
Walschaert gear is made up entirely of comparatively light rods. 
On some of the largest locomotives the Stephenson link motion 
weighs as much as two tons. Two reversals of this large weight 
in each revolution with a fast running locomotive throws a great 
strain upon the engine. 

It is desirable that any valve gear keep its adjustment. In 
this respect the Walschaert is superior to the Stephenson gear. 
The joints in its moving parts are all pin joints with hardened 
bushings which reduce wear to a minimum and which may be 
easily renewed should wear occur. In the Stephenson link motion 
there is a certain amount of sliding between the block and the 
link called the ''slip." The two eccentrics are also subject to 
considerable wear. Both of these parts are under a locomotive 



196 STEAM ENGINES 

and exposed to the dust and grit from the road bed, which causes 
them to wear and cut very fast so that the Stephenson motion 
soon loses its adjustment. 

The Stephenson hnk motion is usually placed between the 
drive wheels of a locomotive. It is therefore not so accessible 
as the Walschaert gear which is placed outside the drive wheels. 
The more accessible location of the valve gear insures its receiving 
better attention from the engineer, especially in the matter of 
lubrication. It might be thought that the more exposed location 
of the Walschaert gear renders it more liable to damage, but 
experience shows that it is no more liable to injury than the 
Stephenson link motion. 

The amount of space occupied by the Stephenson link motion 
under a locomotive has, in some cases, interfered seriously 
with the proper bracing of the frame. The location of the 



Fig. 125. 

Walschaert gear outside the drive wheels leaves the space under 
the locomotive free for proper bracing, and it does not interfere 
with the design of any other part of the locomotive. 

The varying lead of the Stephenson link motion is urged as 
an objection by some engineers who prefer the constant lead of 
the Walschaert gear. 

Woolf Reversing Gear. — This is a simple form of single eccen- 
tric valve gear used on a great many traction and similar engines. 
This form of valve mechanism is illustrated in Fig. 125. The 
eccentric is set 180° from the crank. The eccentric strap carries 
a projecting arm which extends upward a little in front of and 
inclined to a vertical line through the center of the shaft. The 
outer end of this arm carries a block which slides in a straight 
but inclined guide. The eccentric rod is attached to the eccentric 
arm a short distance below the block, and extends in a direction 
almost at right angles to the eccentric arm. The other end of 



REVERSING MECHANISMS 



197 



the eccentric rod is connected to a rocker arm to which the valve 
stem is attached. 

The action of the valve gear may be seen more clearly from 
the center-line diagram shown in Fig. 126. The block Q on the 
upper end of the eccentric arm moves in a straight inclined line in 
the guides while the lower end moves in a circle. This makes 
the point P at which the eccentric rod is attached move in an 
ellipse. The engine is reversed by throwing the guide over to 
the dotted position. The point P then follows the path of the 
dotted ellipse. The point of cut-off may be varied by changing 







< \ 


^ 


RtvcRSE Roc 




.'-^ 


'^'Z)p-A 




Valve Stem 


^j,,,.c,,aBi^32^ 


\/ 


^\j \ 


, 






/ \ 


A 


\ 


- 




\ t^^y^ _ ' 


\ 


Piston Path 






11 
















^^^^'^^^Zio^.^^o^ 


-~y 


^^^^..y' 




Fig. 126. 









the inclination of the guide, and the reverse lever is supplied 
with a notched sector for this purpose. 

The Woolf reversing gear gives a constant lead for all degrees 
of cut-off but the leads for the two ends of the cylinder are not 
necessarily equal. The steam distribution is not the same for 
both directions but the mechanism is usually designed to give 
equal cut-off for both ends of the cylinder. Placing the center 
of the guides a little forward of the center of the shaft makes the 
steam distribution more nearly alike for both directions of 
rotation. 

The simplicity of this reversing gear, together with the fact 
that it gives a fairly good steam distribution, has made it very 
popular for traction and thrashing engines where the engineer is 
usually not an expert. 



CHAPTER XV 
CORLISS VALVE GEARS 

Advantages of the Corliss Valve. — The Corliss valve mechan- 
ism has been described in a previous chapter and the two admis- 
sion and two exhaust valves shown. The use of four valves on 
a steam engine has three distinct advantages. They may be 
located close to the cylinder bore and long ports thus avoided. 
This reduces the clearance volume and the surfaces upon which 
steam may be condensed. When a single valve is used to control 
both admission and exhaust, it is first heated by the live steam 
passing it and then cooled by the exhaust steam passing it. 
This promotes condensation of steam upon the surfaces of the 
cooled valve. The use of separate valves for admission and 
exhaust reduces this condensation. The use of four valves has 
the further advantage that each one may be adjusted separately, 
which simplifies the operation of setting them. 

Besides the advantages which arise from the separation of the 
admission and exhaust functions the Corliss mechanism gives 
an almost ideal motion to the valves. The valves move quickly 
while they are opening and closing thus giving sharp and well- 
defined events of admission, cut-off, release, and compression 
and also avoiding wire-drawing during admission. The valves 
have but little motion after they are closed, when motion is not 
needed, and the motion at this time is slow. This reduces the 
friction of the valve and makes good lubrication possible. 

The slow motion of the valves after they are closed is due to 
the angle at which the valve rods are connected to the wrist plate. 
This angle is such that the valves are given a rapid motion 
when the wrist plate is near its central position and a slower 
motion as the wrist plate approaches either extreme position. 
Fig. 127 will make this clear. The heavy line CC shows the 
position of the steam valve rod when the wrist plate is in its 
extreme position towards the left and the steam hook picks 
up the valve. As the wrist plate moves to the right through the 
angle from C io B the steam arm moves through the angle 

198 



CORLISS VALVE GEARS 



199 



between C and B\ The valve remains closed during this period. 
A further movement of the wrist plate through the angle between 
B and A moves the steam arm through the angle between B' and 
A' and the valve is opened during this period. It will be observed 
that although the wrist plate moves through a smaller angle 
while the valve is open than it does while the valve is closed that 
the steam arm (and also the valve) moves through a larger angle 
while the valve is open than while it is closed. Hence the valve 




Fig. 127. 

has a slow motion while it is closed and a fast motion while it is 
open. This effect is even more pronounced since the wrist 
plate moves slower when near its extreme position (when the 
valve is closed) than it does when near its central position. 

The above remarks apply only to the opening of the steam 
valve, but the same effect is produced by the dashpot in closing 
the valve. As soon as the steam valve is released the dashpot 
moves rapidly and closes the valve. When the dashpot piston 

20 



200 



STEAM ENGINES 



is near the end of its downward stroke, it moves more slowly in 
seating gradually and the valve, which is closed by this time, 
also moves slowly. 

The exhaust valves are connected to the wrist plate at all 
times but they also have a slower motion when closed than when 
opening. It will be observed that the wrist plate moves through 
the small angle between F and E while the exhaust valve arm 
moves through the large angle between F' and E\ The exhaust 
valve is open during this period. While the wrist plate moves 
through the large angle between E and D the exhaust valve arm 
moves through the small angle between £" and D\ the valve being 
closed during this period. Moreover, the motion of the wrist 




Fig. 128. 

plate is slower while moving from E and D than while moving 
from F and E. Each valve rod is usually connected to the wrist 
plate in such position that a line through its axis, if extended, 
would pass through the center of the wrist plate when it is in its 
extreme position. 

Single Eccentric Valve Gear. — The valve gear of some Corliss 
engines is operated by a single eccentric while the valve gear of 
others is operated by two eccentrics. With a single eccentric 
the cut-off may occur anywhere between the beginning and 
35 per cent, to 40 per cent, of the stroke. With two eccentrics, 
one to operate the admission valves and the other to operate the 
exhaust valves, the range of cut-off may be extended considerably 
beyond mid-stroke. 

With a single eccentric valve gear the admission valve must 
be unhooked for cut-off before the eccentric reaches the end of its 



CORLISS VALVE GEARS 201 

throw. When the eccentric reaches the end of its throw it 
begins to move in the opposite direction. If the steam valve, 
which is hooked up at the beginning of the piston stroke, has 
not been released by this time it will not be released at all but 
will be closed by the gradual motion of the eccentric on its return 
stroke. Since the eccentric is set a little more than 90° ahead of 
the crank it will reach the end of its stroke before the crank 
reaches a vertical position, representing mid-stroke of the piston. 
This will be made clear by referring to Fig. 128. 

The admission valve is hooked up when the eccentric is in 
the position OC, at the beginning of its stroke. When the crank 
reaches its dead center position the eccentric is in the position OE 




and the valve has been opened an amount equal to the lead. The 
valve may be tripped by the knock-off cam at any time while the 
eccentric is moving through the angle between E and E\ Dur- 
ing this time the crank is moving from the position OC to the 
position 0C\ This is somewhat less than half the piston stroke. 
If the steam valve is not tripped by the time the eccentric reaches 
the position OE' (and the crank reaches the position OC) it 
will not be tripped at all, but will remain hooked up and it will 
then be closed by gradual motion of the eccentric on its return 
stroke. 

By referring to Fig. 128 it will be seen that if the angle of 
advance is increased the range of cut-off will be decreased, hence 
it might be thought that a greater range of cut-off could be 



202 



STEAM ENGINES 



secured by decreasing the angle of advance. The proper working 
of the exhaust valves which are connected with the eccentric at 
all times, however, sets a limit to the decrease in the angle of 
advance. 

The exhaust valves, being connected with the eccentric at 
all times, have the same motion as a plain slide valve and this mo- 
tion may be shown by a Zeuner valve diagram. Fig. 129 shows a 
Zeuner valve diagram for the head end exhaust valve of a Corliss 
engine, the eccentric being set with the angle of advance AOB. 
Release occurs when the crank is in the position OR, the piston 
being near the end of its stroke from left to right. Compression 
occurs when the crank is in the position OK, the piston being near 
the end of its stroke from right to left. The Zeuner valve diagram 
does not show the position of the cut-off for a Corliss valve gear. 




Fig. 130. 



hence it is of but little use in setting these valves. Corliss valves 
are usually set by measurement and the setting then checked 
from an indicator diagram. 

Setting Corliss Valves. — When the various parts of a Corliss 
valve gear are in proper adjustment the reach rod and the 
eccentric rod should be of such length that both the rocker arm 
and the wrist plate will be plumb when the eccentric is vertical, 
as shown in Fig. 130. 

Since it is difficult to judge by the eye when an eccentric 
is vertical the following method should be used for finding exactly 
its vertical position. A tram shaped as shown at A in Fig. 131 
is made of sheet iron and a hole bored in it large enough to receive 
a scratch awl or pointed nail. With the crotch of the tram placed 
against the shaft and with a scratch awl in the hole in the tram 



CORLISS VALVE GEARS 



203 



the arcs BC and EF are drawn on the eccentric ending at the 
points B and E at the edge of the eccentric. It is convenient 
to rub chalk on the eccentric and the arcs drawn in the chalk 
in order to make them more easily seen. With a pair of dividers 
draw arcs from B and E which meet exactly on the edge of the 
eccentric as at G. By a similar method the point / exactl}^ 
opposite G is located. The point G is furthest away from and 
the point / is nearest the center of the shaft, therefore the line 
IG represents the position of the eccentric. 

The joint in the eccentric strap is usually at right angles to 
the eccentric rod. hence the eccentric may be placed in a 
vertical position by turning it until the point G comes to the 




Fig. 131, 



joint. In case the joint in the eccentric strap is not at right 
angles to the eccentric rod, the following method of placing the 
eccentric in a vertical position may be used. 

Another tram, shaped as shown at H, Fig. 132, is made of steel 
wire. A punch mark J is made on the side of the eccentric rod 
and on its center line. With one end of the tram in the mark J 
arcs are drawn on the eccentric strap ending at the points K and 
L on the edge of the strap. With a pair of dividers draw arcs 
M from K and L which meet exactly on the edge of the strap. 
When the point M coincides with either the points G or I the 
eccentric is exactly on center. 

With the dividers set to the length KM, strike arcs P and R 
from / which end at the edge of the strap. Draw arcs S from P 



204 



STEAM ENGINES 



and K which meet at the edge of the strap and arcs T from R and 
L which meet at the edge of the strap. When the point G coin- 
cides with either the points S or T the eccentric is exactly vertical. 




Fig. 132. 



The manufacturers of Corliss engines usually place a mark A, 
Fig. 133, on the hub of the wrist plate and three marks D, B, and 
C on the wrist plate stud. When A coincides with B the wrist 



B 



c 



_z 



■^ 



7^~^ 




D 



Fig. 133. 



plate is in its central position, and in this position the rocker 
arm and eccentric should be vertical, and if they are not vertical 
the length of the eccentric and reach rods should be adjusted 



CORLISS VALVE GEARS 



205 



until they are vertical. When A coincides with D the wrist 
plate is at one extreme of its travel and the eccentric is on dead 
center; and when A coincides with C the wrist plate is at the 
other extreme of its travel and the eccentric is on the other dead 
center. If the marks A, B, C, and D are not on the wrist plate 
and stud they should be placed on with a chisel, A and B being 
marked when the eccentric is vertical and the rocker arm and 
wrist plate are plumb, and D and C being marked opposite A 
when the eccentric is on its dead centers. 

The reach rod is now disconnected from the wrist plate and the 
wrist plate placed on its central position (so that A coincides with 
B, Fig. 133). With the wrist plate in its central position and both 




Fig. 134. 



steam valves hooked on, the valves should have the proper lap. 
The amount of the lap may be measured by removing the steam 
bonnets on the side of the cylinder opposite the wrist plate and 
inspecting the marks on the valve and cylinder. These marks, 
as shown in Fig. 134, are placed on by the manufacturer of the 
engine, the mark F on the valve being opposite its working edge, 
and E, on the cyHnder being opposite the edge of the port. When 
F coincides with E the edge of the valve is exactly in line with 
the edge of the port. By removing the valve it can be deter- 
mined which edge of the port the mark E on the cylinder is 
opposite and therefore it will be known on which side of E the 
mark F should be for the valve to have lap. The lap is measured 
with a pair of dividers, being the distance between E and F, 



206 



STEAM ENGINES 



Fig. 134. With the wrist plate in its central position the laps 
on the two ends of the cylinder should be equal. The laps may 
be adjusted by lengthening or shortening the radial rods, which 
are provided with left- and right-hand thread connections for 
this purpose. The proper amount of lap to give the steam valves 
depends upon the size of cylinder and may be determined from 
the following table: 



Table for Setting Corliss Valves 



Diam. of cylinder, 


Lap of steam valves, 


Lap of exhaust 


Lead of steam 


inches 


inches 


valves, inches 


valves 


8 


He 


He 


Vzi 


10 


He 


He 


/^2 


12 


He 


He 


Yzi 


14 


^ 


Vs 


H2 


16 


M 


% 


H2 


18 


M 


Vb 


ys2 


20 


M 


Vs 


ys2 


22 


He 


He 


%4 . 


24 


He 


He 


%4 


26 


He 


He 


%4 


28 


He 


He 


%4 


30 


He 


He 


%4 


32 


% 


y. 


He 


34 


% 


M 


He 


36 


H 


M 


He 



The exhaust valves are now given equal laps by adjusting the 
length of their radial rods in the same manner as was done with 
the steam valves. Marks are placed by the manufacturer on the 
cylinder and on exhaust valves, similar to those on the steam 
valves, and the exhaust laps may be set by measurement between 
these marks. The amount of the exhaust lap is determined from 
the above table. 

The wrist plate is now turned to its head end extreme position 
and the length of the dashpot rod adjusted so there will be equal 
clearance around the catch block as shown in Fig. 135, at G and 
H. It is very important to adjust the length of the dashpot 
rods properly because if they are made too short the valves will 
not hook on and if they are too long the valve stem is liable to be 
bent or the steam bracket broken, or both. Now turn the wrist 
plate to its crank end extreme position and adjust the length of 



CORLISS VALVE GEARS 



207 



the crank end dashpot rod as was done for the head end dashpot 
rod. 

The steam valves may now be given their proper lead, as in- 
dicated by the above table. To do this the engine is placed on 
its head end dead center, using a tram to locate the dead center 
exactly. The eccentric is then loosened on the shaft and the 
reach rod hooked to the wrist plate. The wrist plate is then 
moved over to its head end extreme position in order to hook 
up the head end steam valve. The eccentric is then turned 
around on the shaft until the port is open the amount of the 




Fig. 135. 



desired lead. The eccentric is then fastened in this position. 
The port opening may be measured by the lines on the valve and 
back side of the cylinder. The engine should now be turned 
to the crank end dead center, the crank end steam valves hooked 
up, and the lead measured to see if it is equal to that on the head 
end. If it is not, any slight adjustment that may be required 
can be made by moving the eccentric. 

The governor and governor rods should next be a:' is^ed if 
they require it. Fig. 136 shows a common form of Corliss engine 



208 



STEAM ENGINES 



governor with its connections to the valves, parts of the governor 
being cut away to show its construction. As shown here the 
parts of the governor are in the position they will occupy when 




<//c 


e, ^ 


WL 


1 1 


\J/ * ^« 




Ik l^ii 




JPk ^ '^ -^ 










^^^^/^^'^ ^ 




gg^^y^ 





CD 
CO 



M 



the engine is running at its normal speed. If the speed rises 
above normal, centrifugal force throws the governor balls 
farther from the center. This raises the cross bar and with it 
the drop rod arm and throws the knock-off levers (on the valve 



CORLISS VALVE GEARS 209 

stems) around so that the knock-off cams strike the steam hook 
earher and thus cause an earher cut-off. If the speed falls below 
normal the knock-off cams are moved in the opposite direction 
and cut-off occurs later. 

If the belt which runs the governor should break, the cross bar 
would drop to its lowest position and this would make cut-off 
come at the latest possible point in the stroke or the steam hooks 
would not be disengaged at all and the cylinder would take steam 
for the entire stroke. This would cause the engine to run 
away. 

In order to prevent the engine from running away if the 
governor belt breaks or any other accident happens which would 
throw the governor to its lowest position, safety cams are placed 
on the knock-off levers. When the governor falls to its lowest 
position the knock-off levers are thrown around far enough to 
bring the safety cams under the steam hooks, thus preventing 
the admission of any steam to the cylinder. If the safety cams 
are allowed to come into action as described above the engine 
could not be started, after it was shut down, until the governor 
was raised far enough to prevent the safety cams from coming 
in contact with the steam hooks. It would be a nuisance to have 
to raise the governor every time the engine is started, hence a 
governor safety stop is placed on the governor to prevent it 
from falling to its lowest position and bringing the safety cams 
into action when the engine is shut down by hand. 

When the engine is to be shut down by hand the governor 
safety stop is raised to the position shown in Fig. 136. The cross 
bar will then rest on the safety stop, which is high enough above 
its lowest position to prevent the safety cams from coming into 
operation. In this position of the cross bar the valves will hook 
up and open when steam is turned on again. As soon as the 
engine is started again the safety stop falls to one side. If, then, 
the governor belt should break the cross bar would fall to its 
lowest position and bring the safety cams into operation. 

When the engine is running at normal speed the cross bar on 
the governor is about halfway between the upper limit of its 
travel and the end of the safety stop (in its raised position), 
hence in adjusting the governor it should be blocked up to this 
position. The length of the governor drop rod is then adjusted 
until the drop rod arm is horizontal and the bell crank stands 
vertically. The governor is then unblocked, the engine started 



210 STEAM ENGINES 

slowly, and the length of the governor rods adjusted so that 
cut-off is equal on both ends of the cyKnder. The governor rods 
are provided with right- and left-hand screws so their length may 
be changed without stopping the engine. An indicator diagram 
should be used to determine when the cut-off is equal on the two 
ends, as well as for all other valve adjustments. 



CHAPTER XVI 
GOVERNING 

Governing. — The work that most steam engines do requires a 
constant or practically constant speed of rotation. This require- 
ment is more difficult to meet than might at first appear, and 
much thought has been expended on this problem in order to 
solve it satisfactorily. 

As mentioned in a previous chapter changes of speed occur in 
two entirely different ways. First, the speed may change during 
a single revolution or cycle of the engine due to a variation in 
pressure against the piston and to the angle at which the force 
of the steam pressure is transmitted to the crank. Second, 
the speed may change due to a change of load or to varying 
boiler pressure, such a change extending over a period of more 
than one revolution. The first of these kinds of speed vari^ltion 
is taken care of by the flywheel which stores up energy during 
one part of a revolution and gives it out again during another 
part, as explained in a previous chapter. The second kind of 
speed variation must, however, be corrected by some kind of 
controlling device, or governor. 

If the load on an engine is increased or if the boiler pressure 
of the steam becomes less the speed of the engine will decrease. 
On the other hand, if the load decreases or the boiler pressure in- 
creases the speed of the engine will increase. The governor is for 
the purpose of regulating the supply of steam to the engine so that 
its speed will remain constant or practically constant. In order 
to do this a steam engine governor either operates on a throttle 
valve placed between the engine and boiler to change the pressure 
of the steam which is being admitted to the engine; or it alters 
the volume of steam admitted to the engine by changing the 
point of cut-off. 

Whatever method of controlling the speed is used, no governor 
can control the speed perfectly because the governor is run by the 
engine itself and some change in speed must occur before the 

211 



212 STEAM ENGINES 

governor can operate. That is, the governor operates on account 
of a change of speed, hence the governor cannot keep the speed 
of the engine absolutely constant. Also, any change of speed 
which occurs after the steam has passed beyond the influence of 
the governor cannot be controlled until the next stroke of the pis- 
ton. For example, if the load changes after cut-off has occurred 
this may affect the speed of the engine and the governor can- 
not have any effect because it can only operate on the next 
admission of steam or during the next stroke of the piston. This 
is especially true of compound engines, where the governor 
controls the supply of steam to the high pressure cylinder only. 
If a change of load occurs after cut-off in the high pressure cylin- 
der, the steam in this cylinder expands and does work in the cylin- 
der and then passes into the low pressure cylinder and again 
expands and does work, all outside of the control of the governor, 
which can only act upon the next admission to the high pressure 
cylinder. But for all of the difficulties in the way of securing 
close speed regulation, a good steam engine governor will control 
the speed within 2 per cent, of its normal speed. 

Pendulum Governor. — Nearly all steam engine governors 
operate through centrifugal force. They usually consist of a 
pair of weights revolving about a spindle which is driven from the 
engine shaft. The centrifugal force of the revolving weights is 
resisted by some controlling force, such as gravity, the tension 
of a spring, or both. When the engine (and governor) is running 
at constant speed the weights take up a fixed position at which the 
controlling force just balances the centrifugal force. When an 
increase of speed occurs the additional centrifugal force causes 
the weights to move outward to a new position, and in moving 
outward they act upon the throttle valve or some form of auto- 
matic gear by which the cut-off is varied so that the speed is 
reduced. 

The most common forms of centrifugal governors are those 
known as pendulum governors. In these governors the spindle is 
vertical and there are two weights, each of which is placed at the 
end of an arm. The two arms are suspended from the top of the 
spindle and pivoted at or near it. When the spindle is rotated 
the weights move outward and upward and their upward motion 
is resisted by the force of gravity. When the engine is running 
at constant speed the weights take up a position in which the 
force of gravity, or other controlling force, just balances the cen- 



GOVERNING 



213 



trifugal force. If the speed of the engine is increased or decreased 
the weights take up a new position in which the controlling 
force balances the centrifugal force developed by the revolving 
weights. 

A form of governor commonly used with plain slide valve 
engines and operating on the above principles is illustrated in 




Fig. 137. 



Fig. 137. This is a throttling governor since it acts upon a 
throttle valve and thus controls the pressure of the steam ad- 
mitted to the cylinder. In the governor illustrated here the 
opening A is connected to the steam chest and the opening B 
connected to the steam pipe leading from the boiler, so that the 
steam passes through the valve C before entering the cylinder. 
The governor is run by a belt from the engine shaft to the pulley 



214 STEAM ENGINES 

E. This transmits motion through the gear wheels D and F 
to the vertical spindle to which the weights G and G are attached. 
As the speed of the engine increases the weights move outward 
and upward and press downward upon the valve stem H thus 
partly closing the valve C. In the same way, a decrease in the 
speed of the engine causes the weights to assume a lower position 
and the valve stem H rises and admits more steam to the cylinder, 
thus causing an increase of speed. 

The upward movement of the weights is resisted partly by the 
force of gravity and partly by the tension of the spring K, so that 
for any particular speed of the engine the weights will take a 
position at which their centrifugal force is just balanced by the 
force of gravity and the tension of the spring. If the engine 
departs from this speed the weights will r ise or fall until the 
steam pressure is adjusted to suit the load which the engine is 
carrying. 

The speed of the engine can be changed by changing the 
tension of the spring K by means of adjusting screw T. If the 
tension of the spring is increased, the weights will have to revolve 
faster in order to secure a given opening of the throttle valve, 
and thus the speed of the engine will be increased. In the same 
way, the speed of the engine may be reduced by decreasing the 
tension of the spring. 

The form of governor described above was invented by James 
Watt, one of the early inventors of the steam engine, and for this 
reason it is sometimes called the Watt governor. It is one of 
the simplest forms of steam engine governors used at the present 
time. 

Stability. — A governor is said to be stable when there is a 
definite position of the weights for any definite speed; that is, 
if the speed of the engine changes by any amount the weights 
move up or down to a new position which corresponds to that 
particular speed, and then remain in this new position until 
there is another change of speed. From the preceding descrip- 
tion of the Watt governor it will be seen that the new position 
of the weights gives a larger or smaller opening of the throttle 
valve and this serves to bring the speed back to normal. The 
speed is thus automatically maintained at or near the number of 
r.p.m. for which the governor is set. 

If a governor was unstable it would have no definite position 
for a given speed and its movements would be irregular and un- 



GOVERNING 



215 



certain. For this reason it could not maintain a constant speed 
of the engine and would therefore not be suitable for governing 
a steam engine from which a constant speed was desired. 

It is evident from the above discussion that stability is a 
desirable and even necessary quality of a steam engine governor 
because a stable governor is always in equilibrium and exercises 
a positive control over the speed. 

In order for a governor to be stable the controlling force, or 
the force acting against the rise of the weights, must increase at 
a faster rate than the radius of the circle about which the weights 
are revolving, and the larger this ratio between the controlling 
force and the radius of the circle about which the weights are 
revolving, the greater will be the stability of the governor. 





Fig. 138. 



A change in the speed of a governor causes a change in the 
position of the weights, and if the governor is stable there is 
only one position of the weights which correspond with a given 
speed. Since the steam supply depends upon the position of 
the weights, a stable governor cannot maintain a strictly con- 
stant speed, because if the boiler pressure or load changes a cer- 
tain displacement of the weights s necessary to admit more or 
less steam and the weights can maintain this new position only 
by turning faster or slower. However, the variaton from a 
constant speed can be reduced by reducing the stability of the 
governor. 

The ordinary forms of pendulum governors such as illustrated 
in Fig. 138 are stable and also the crossed arm form of pendulum 
governor illustrated in Fig. 139 provided the points of support 
are located close to the central column. 

Sensibility. — The movement of the weights of a governor from 
their lowest to their highest position can produce only a certain 

21 



216 



STEAM ENGINES 



movement of the regulating mechanism, whether it is a throttle 
valve or a cut-off attachment. This will produce the greatest 
change of speed for which the governor is responsible. If an 
engine is overloaded or if the steam pressure is too low, the speed 
may drop even after the governor has done all that it can do to 
admit steam freely, but the variation in speed for which the 
governor is responsible is only that which will cause the weights 
to move from the position of no steam to the position of full 
steam. When a small variation of speed is sufficient to do this 
the governor is said to be sensitive. 

It is evident from the above discussion that the more sensitive 
a governor is the less stable it must be. As both of these features 
can be controlled in the design of the governing mechanism, the 




Fig. 139. 



designer aims at securing a governor which is stable and which at 
the same time is sensitive enough to control the speed within 
the limits needed for the kind of work the engine is to do. 

It is absolutely necessary that a steam engine governor be 
stable and it is highly desirable that it be sensitive. However, 
it should not be too sensitive as this causes the engine to hunt 
or over-govern. Hunting is brought about by the conditions 
mentioned below. 

When an alteration of speed begins, the governor does not act 
immediately because the governor can only operate after a 
change of speed has occurred. Moreover, a change in position 
of the governor does not affect the speed of the engine immedi- 
ately both on account of the inertia of the moving parts of the 
engine which has the effect of resisting a change of speed, and 
also because of the energy contained in the steam which has 
passed the control of the governor. If the governor is of the 



GOVERNING 



217 



throttling type, the steam chest is filled with steam which has 
passed the control of the governor at the time the change of 
speed begins, and if the governor acts upon the cut-off its oppor- 
tunity for controlling the speed has passed if cut-off has occurred. 
Hence, there is a certain time lag between the governor and the 
engine speed. The consequence of this is that, if the governor 
is too sensitive, by the time the change in engine speed has had 




Fig. 140. 



full effect upon the governor, it is forced into a position of over- 
control or a position beyond that which is necessary to bring 
the engine speed back to normal. The speed of the engine then 
begins to change in the opposite direction and, for the same 
reasons, the governor is forced into a position of over-control 
in the opposite direction. Thus a state of forced oscillation 
is set up which causes the speed to be first too high and then too 
low, a condition known as hunting. 



218 



STEAM ENGINES 



Hunting is avoided by allowing a certain margin of stability 
in the governor, that is, by not making it too sensitive, and also by 
the use of dashpots attached to the governor in such way as to 
dampen its motion in case it is too sensitive. 

Some governors, on account of their form, are much more 
sensitive than others. It has been found that if the form of 
governor is such that the weights, in rising, follow a parabola 
instead of a circle the governor will be extremely sensitive. A 
governor constructed in this way, illustrated in Fig. 140, is so 
sensitive that an air cylinder and piston is placed at the top of 




Fig. 141. 



the column to check the movement of the weights. Other forms 
of sensitive governors are the Proll governor illustrated in Fig. 141 
and the Hartnell governor illustrated in Fig. 142. In the Hart- 
nell governor, the weights move in a practically horizontal path 
and the controlling force is furnished by a coil spring shown at 
the top of the central column, and the sensitiveness of the 
governor can be adjusted by means of this spring. This form of 
governor is more suitable for high speed engines than for low 
speed ones on account of the small size of the weights. 

The Proll governor illustrated in Fig. 141 is better adapted to 
slow speed engines, such as the Corliss engine. This type of 
governor is known as a loaded governor on account of the heavy 



GOVERNING 



219 



weight placed around the central column. This weight revolves 
with the governor and has the effect of increasing the controlling 
force without adding to the centrifugal force, as would be the 
case if the additional weight was placed at the ends of the rotating 
arms. 

The advantages of a loaded governor are that it is more 
powerful than an unloaded one, that the increase in power is 
gained without a corresponding loss of sensitiveness, and that it 




Fig. 142. 



may be run at a lower speed than an unloaded governor having 
balls of equal size. 

A powerful governor is necessary in order to overcome the 
friction of the moving parts of the governor and controlling 
mechanism. Friction in a governor and its connected mechanism 
has the effect of increasing the controlling force and thus reducing 
the sensitiveness of the governor. If the controlling force of a 
governor, neglecting friction, is represented by F and the force 
necessary to overcome friction is represented by / then for an 
increase of speed the centrifugal force acting on the weights 
must be increased to F + / in order to change the position of the 



220 



STEAM ENGINES 



governor and also the centrifugal force must be decreased to F — 
/ in order to change the position of the governor for a decreasing 
speed. 

A loaded governor of the type shown in Fig. 141 or Fig. 136, 
that is, a governor having four arms, has the further advantage 
that the vertical movement of the collar, or central column, 
is twice as great as the movement of the weights at the ends of 
the rotating arms. That is, for a given change of speed a gover- 
nor of the types shown in Figs. 136 and 141 produces twice as 
much motion in the controlling mechanism as in the plain pendu- 
lum governor such as illustrated in Fig. 138a. The result of 




Fig. 143. 



this is that these governors may be run at a less speed than the 
Watt governor, hence their general use on low speed engines of 
the Corliss type. 

Shaft Governors. — Governors which are located in the fly- 
wheel and which turn with the flywheel are commonly called 
shaft governors. These are usually attached to the spokes of 
the flywheel and operate by shifting the position of the eccentric 
and thus changing the point of cut-off so the amount of steam 
admitted to the cylinder is in proportion to the load which the 
engine carries. 



GOVERNING 



221 



One example of shaft governor operating on this principle is 
described in Chapter 13 and illustrated in Fig. 108. The prin- 
ciples upon which a governor of this type operates are the same 
as those upon which the centrifugal pendulum governor operates. 
The similarity of operation of these two kinds of governors may 
be seen by a study of Fig. 143. An arm is pivoted at some point 
on the flywheel and to the end of the arm is attached a weight W, 
the length of the arm being A. At a distance A' from the pivot a 
spring is attached to the arm and is arranged so as to act at 
practically right angles to the arm. The centrifugal force of 
the weight will be balanced by the pull of the spring in the same 
manner as gravity balances the centrifugal force in a pendulum 
governor. 





B 



Fig. 144. 



With a shaft governor, the speed at which the engine will run 
increases as the tension in the spring is increased, or as the 
distance between the pivot and the point of attachment of the 
spring is increased, or as the weight at the end of the arm is made 
less, because in these cases the governor must rotate faster to 
maintain the same position. As it is usually impractical to 
change the weight of the arms, the speed is usually changed by 
changing the tension of the spring. For this purpose the spring 
is provided with a turnbuckle or some other arrangement by 
which its tension may be readily changed. 

It is sometimes desired to reverse the direction of rotation 
of an engine which is fitted with a shaft governor. This can be 



222 



STEAM ENGINES 



done with some engines but with others it cannot be done as the 
manufacturer has not made provision for it. 

In order to reverse the direction of rotation of a shde valve 
engine the eccentric must be turned through an angle of 180 
degrees on the shaft. In a shaft governor the eccentric is con- 
nected directly to the arms of the governor, consequently the 
arms must be turned around so as to swing in the opposite 
direction and the attachment of the springs to the rim or spokes 
of the flywheel reversed by attaching them to other holes, if 
these have been provided by the manufacturer. The changes 
required in order to reverse the direction of rotation of an engine 
supplied with a shaft governor are illustrated in Fig. 144. In 
this illustration (a) shows the arrangement of the parts for one 





Fig. 145. 



direction of rotation and (6) shows the arrangement of the parts 
for the opposite direction of rotation. As each manufacturer 
of high speed engines has his own arrangement of shaft governor 
it is impossible to give definite directions for the proper arrange- 
ment of the governor in order to reverse the direction of rotation, 
consequently, if this is desired, it is best to write to the manu- 
facturer before attempting to make any changes in the governor. 
Inertia Governor.^ — ^A form of shaft governor invented by F. 
M. Rites and known as the Rites inertia governor is used on sev- 
eral makes of automatic high speed engine. In this governor, 
which is illustrated in Fig. 145, a heavy bar on .the flywheel, 
carrying two weights F and A, swings about the pin B. The 
eccentric, which is usually only a pin located near the center of 



GOVERNING 223 

the engine shaft is carried by the arm. As the arm swings about 
the pin B, the eccentric pin C swings closer to or further away 
from the center of the shaft and thus changes the eccentricity. 
The controlhng force is furnished by the coil spring D, which has 
one end fastened to the arm of the governor and the other end 
fastened to a spoke of the flywheel. 

This governor operates by the force of inertia, or the tendency 
of the weights to keep on moving at a constant speed, when the 
speed of the flywheel changes. If the engine is running at a 
constant speed the flywheel and governor weights will be turning 
at the same rate. Referring to Fig. 145, suppose a load is sud- 
denly put on the engine. This slackens the speed of the flywheel, 
but the inertia of the governor weights causes them to move 
forward at the same rate as before. This moves the eccentric 
pin further away from the center of the shaft, which increases 
the eccentricity and causes a later cut-off. Also, if the load 
should be decreased, the speed of the flywheel will increase, 
causing the governor weights to lag behind and reduce the 
eccentricity. This causes cut-off to occur earlier and bring the 
speed back to its normal value. 

The inertia governor described above is extremely simple but 
in securing this simplicity some things have been sacrificed. One 
of these is that the governor does not give a constant lead, which 
is desirable for a constant speed engine. It will be seen also 
that the governor is unbalanced, since it is pivoted away from the 
center of the shaft. This causes the arm to tend to fall forward 
during one-half of the revolution and to fall backward during the 
other half of the revolution. If the speed is high, say over 250 
revolutions per minute, this effect is not noticeable, but for 
lower speeds it will affect the cut-off. If the speed is reduced 
much below 200 revolutions per minute, this unbalancing 
effect becomes noticeable as a jerk in the governor action 
which may send the governor arm through its whole range 
every second or third revolution. 

To prevent this action, a drag or brake spring is attached to 
the rim of the flywheel in such manner as to bear against one of 
the weights with enough force to prevent sudden swinging but 
not enough to prevent the governor from swinging when there is 
a change in load. In addition to this dampening spring there is 
also a spring bumper fastened to the rim of the flywheel to prevent 
the arm from swinging too far and damaging the valve. 

22 



224 STEAM ENGINES 

In some forms of inertia governor a second arm is placed 
parallel with the one carrying the weights and arranged so the 
whole governor will be balanced. This makes the governor more 
complicated but makes it suitable for running at low speed and 
it gives the same sensitiveness as the unbalanced governor at 
high speeds. 



CHAPTER XVII 
COMPOUND ENGINES 

Compounding. — The low efficiency of the steam engine shows 
that a large part of the heat energy supplied to it is not turned 
into useful work, but is lost or wasted. Even the best engines 
utilize only about 20 per cent, of the heat supplied to them, 
leaving about 80 per cent, to be accounted for by the various 
losses incident to the operation of the engine. Radiation of 
heat from the engine and the friction of the moving parts account 
for only a small part of the loss. A much larger part is accounted 
for by the heat contained in the exhaust steam. The loss from 
this source may be reduced considerably by the use of a conden- 
ser, which lowers the exhaust pressure and makes a larger propor- 
tion of the total supply of heat available for useful work; but, 
even with the use of a condenser, the loss of heat in the exhaust 
is considerable. 

For a long time after the steam engine was invented, the 
three sources of loss mentioned above, namely, radiation, friction, 
and loss of heat in the exhaust were thought to be the only ones. 
It was discovered later, however, that another serious source of 
loss comes from the interchange of heat between the steam and 
the cylinder walls, which results in condensation of steam inside the 
cylinder. The manner in which cylinder condensation produces 
a loss has been fully discussed in Chapter 8 and the student 
should review this chapter at this time in order to understand 
more fully the principles underlying the compound engine. 

Since cylinder condensation produces such large losses in the 
operation of steam engines, it becomes a matter of considerable 
importance to understand the causes of cylinder condensation 
and the means employed for reducing it. The principal cause of 
cylinder condensation is the large range of temperature to which 
the walls of the cylinder (including head and piston) are subjected 
during each revolution of the engine. This range of temperature 
is due to the expansion of the steam in the cylinder from the high 
pressure of admission to the relatively low pressure of the exhaust. 
23 225 



226 STEAM ENGINES 

The variation of pressure, and therefore the range of tempera- 
ture, in an engine cyUnder depends upon the cut-off. With a 
fixed exhaust pressure an early cut-off will produce a large varia- 
tion of pressure during expansion and a late cut-off will produce a 
small variation of pressure. It is evident that an early cut-off 
is necessary to the economical use of the steam because this 
permits the expansive force of the steam to be utilized more 
fully than with a late cut-off and small expansion. The engine 
designer is therefore confronted with two opposing conditions. 
On the one hand, an early cut-off increases the losses from cylin- 
der condensation, and on the other hand, an early cut-off is 
necessary if the steam is to be expanded through its full range and 
utilized efficiently. 

One of the means most commonly employed for reducing the 
losses from cylinder condensation is to divide the total expansion 
of the steam into two or more parts and to perform each part of 
the expansion in a separate cylinder, thereby reducing the range 
of temperature in each cylinder. This is called compounding, 
and engines in which the total expansion of the steam is divided 
between two cylinders are called compound engines. 

Since the losses from cylinder condensation increase as the 
total range of pressure through which the steam is expanded 
increases, the number of parts into which the total expansion 
should be divided, in compounding, depends upon the pressure 
of the steam supplied to the engine. It also depends, to a certain 
extent, upon the kind of work for which the engine is intended. 
In marine work, where compounding is more generally practised 
than in stationary work, the number of parts into which the 
total expansion is divided for different boiler or admission pres- 
sures is about as follows: 

Simple engines 30 to 70 lb. per sq. in. gage 

Compound 80 to 120 lb. per sq. in. gage 

Triple expansion 140 to 180 lb. per sq. in. gage 

Quadruple expansion 200 to 250 lb. per sq. in. gage 

In stationary work there is a tendency to divide the total 
expansion of the steam into a fewer number of parts and to use 
higher pressures. Compound condensing engines are often 
run with pressures of 120 to 150 lb. per sq. in. gage, while the 
compound locomotive, which is not used with a condenser, is 



COMPOUND ENGINES 227 

sometimes supplied with steam having a pressure of 200 to 225 
lb. per sq. in. gage. 

Expansion of Steam.^ — An ideal expansion line for steam 
expanding from 120 lb. per sq. in. absolute pressure to an exhaust 
pressure of 1.6 lb. per sq. in. absolute pressure is shown in Fig. 
146. The diagram ABCDEFG represents an ideal indicator 
diagram if the total expansion of the steam occurred in a single 
cylinder having no clearance. The line AB represents the admis- 
sion line, which is very short compared with the total length 
of the stroke, which is represented by the line GF, and which 



I^O 


B p s 120 Lb 




IZO " 




Its 34l« 




100- 




\ 




N 




1 




80- 




\ 




* 




\ 




€0- 


: 


\ 




-40- 




\. 




20- 


J 




t =* 225* 










a . 


G 


P = 1,6 Lb., t = lie* 


' 1 



Fig. 146. 



also represents the volume of the steam after expansion. The 
line GF therefore also represents the volume of the cylinder. The 
short admission line is necessary if the steam is to be expanded 
through the entire range of pressure in one cylinder. It will 
be observed that if the entire expansion occurred in a single 
cylinder, this cylinder would have to be large enough to accommo- 
date the entire volume of steam GF, and would have to be strong 
enough to withstand the full pressure of 120 lb. per sq. in. The 
objection to this would be that the cost of such a cylinder would 
be excessive; there would be a waste of power in overcoming 
friction; and the large size of the cylinder would expose so much 
surface to the cooling action of the exhaust steam that condensa- 
tion would be excessive. 

Now, suppose that a line JD be drawn across the diagram at 



228 STEAM ENGINES 

such a height that the area of the diagram will be divided into 
two approximately equal parts. If the two parts into which the 
expansion of the steam is divided are performed in separate 
cylinders, the first one will have a volume JC and would be built 
to withstand the full steam pressure of 120 lb. per sq. in. This 
cylinder would admit a volume of steam AB and would expand 
it to the volume JC. The expansion would not be carried further 
than the point C because it is desirable to have enough pressure 
in the cylinder at release to force the steam out of the cylinder 
rapidly, and also because the extra amount of work obtained 
by complete expansion to the point D would not be enough to 
balance the work lost in friction while the piston was moving 
through this part of the stroke. The exhaust from the first 
cylinder would form the supply for the second cylinder. This 
cylinder would have a volume GF, the same as would a cylinder 
designed for the entire expansion, but, as the supply of steam 
for the second cylinder has a pressure of only 19 lb. per sq. in. 
it would not have to be so strong as a single cylinder designed 
for 120 lb. pressure, hence would be cheaper to construct. 

The second cylinder would admit the volume of steam JD at a 
pressure of 19 lb. per sq. in. and would expand it to the volume 
GF, when its pressure would be 1.6 lb. per sq. in. If the total 
expansion occurred in a single cylinder, this cylinder would be 
subjected to the full range of temperature, 223°, and, since its 
wall surface would be large, condensation would be excessive. 
By dividing the expansion into two parts, each cylinder experi- 
ences a range of temperature of only about 112°, that is, the 
range of temperature has been cut in half and the cylinder sur- 
face has not been doubled, hence the condensation and reevapo- 
ration in the two cylinders would be less than in a single cylinder 
subjected to the entire range of temperature. This decreases 
materially the large loss of heat that would otherwise occur 
through condensation and reevaporation; but, on the other 
hand, the engine would be more complicated and therefore more 
expensive, and the friction loss would be increased by the greater 
number of moving parts. 

Compound Engines. — Compound engines are divided into 
two classes, based upon the arrangement of cylinders. These 
are called tandem-compound, in which one cylinder is placed 
behind the other, and cross-compound, in which the cylinders 
are placed parallel with each other. 



COMPOUND ENGINES 



229 



The tandem engine, as illustrated in Fig. 147, has only one 
piston rod, connecting rod, and crank. The piston rod extends 
from one cylinder through the other and has both pistons 




Fig. 147. 



attached to it. The exhaust pipe from the high pressure cylin- 
der passes directly to the low pressure cylinder, and, as this pipe 
is short, it has but little storage capacity; therefore it may be 
considered that the high pressure cylinder exhausts directly into 




Fig. 148. 



the low pressure cylinder. The tandem-compound engine is 
simple in construction, but the parts must be made large in order 
to carry the heavy stresses. 

The cross-compound engine, illustrated in Fig. 148, has two 



230 



STEAM ENGINES 



pistons, piston rods, connecting rods, and cranks, hence it is 
similar to two simple engines placed parallel with each other and 
connected to the same shaft. The cranks are usually placed 
90° apart, which gives a more uniform turning effort on the shaft. 
Since each side of the engine transmits only one-half of the power, 
the parts of the engine are made smaller, but the larger number 
of parts makes this type of engine more expensive than the tan- 
dem-compound. The exhaust pipe from the high-pressure 
cylinder extends across to the low-pressure cylinder and contains 
a receiver or vessel in which steam may be stored. This is made 
necessary by the cranks being placed 90° apart, as explained in a 
later paragraph. 

The action of the steam in the two classes of engines mentioned 




Fig. 149. 

above is quite different. In the tandem engine the pistons 
have the same length of stroke, and move in unison with each 
other, beginning a stroke at the same time and ending it at the 
same time. For this reason the steam exhausted from the high- 
pressure cylinder may be passed directly into the low-pressure 
cylinder without having any valves on the latter cylinder, and 
without any storage space or receiver between the cylinders. 
In this case the valves and governor on the high-pressure cylin- 
der control the action of the steam and the amount of work 
performed in both cylinders. 

Cross -Compound Engines. — The action of the steam in both 
cylinders of a cross-compound engine with cranks set 180° apart 
and without valves on the low-pressure cyUnder is very similar 



COMPOUND ENGINES 231 

to that in a tandem-compound engine, and may be studied best 
by considering the indicator diagrams shown in Fig. 149. This 
illustration shows the diagram from the high-pressure cylinder, 
marked H.P. and that from the low-pressure cylinder, marked 
L.P., placed in their correct relative positions, that is, so that the 
exhaust stroke for the high-pressure cylinder is the admission 
stroke for the low-pressure cylinder. These diagrams do not, 
however, show correctly the division of work between the cylin- 
ders, because, being drawn to the same scale of pressure and 
stroke, they do not take into account the different diameters of 
the cylinders. 

After the supply of steam is cut off from the high-pressure 
cylinder, the steam expands in this cylinder until released. 
During exhaust from the high-pressure cylinder, the steam 
flows directly into the low-pressure cylinder. Since the diameter 
of the low-pressure cylinder is larger than that of the high-pres- 
sure cylinder and both pistons move at the same speed, the 
volume displaced in the low-pressure cylinder is greater than 
that displaced in the high-pressure cylinder. The result of this 
is that each cubic foot of exhaust steam pushed out of the high- 
pressure cylinder flows into a larger volume than one cubic foot 
in the low-pressure, and its pressure therefore falls. This is 
why the exhaust from the high-pressure cylinder and the admis- 
sion to the low-pressure cylinder show a continually falling 
pressure. When the point of compression in the high-pressure 
cylinder is reached, the supply of steam for the low-pressure 
cylinder is stopped and the steam then in the low-pressure cyl- 
inder expands with a rapidly falling pressure, since no new 
steam is being supplied. 

It will be observed from Fig. 149 that the range in temperature 
in the high-pressure cylinder is that represented by the change 
in pressure from A to C, which is greater than it would have been 
if there was less drop in pressure during exhaust. Also the range 
in temperature in the low-pressure cylinder is that due to the 
difference in pressure between E and the exhaust pressure from 
the low-pressure cylinder. Since the pressure at E is greater 
than at C, the range in temperature is greater in both cylinders 
than would be indicated by a division of the work into two equal 
parts. 

The above analysis of the action of steam in the cylinders of a 
cross-compound engine applies only to those engines which have 



232 



STEAM ENGINES 



no valves on the low-pressure cylinder or to those engines which 
have only one valve for both cylinders and this valve so arranged 
that cut-off in the low-pressure cylinder occurs at the same time 
as compression in the high-pressure cylinder. This type of 
engine is not used to a large extent and is made only in compara- 
tively small sizes. A more common arrangement, either in 
tandem-compound engines or in cross-compound engines with 
cranks placed 90° apart, is to have separate valves on each 
cylinder which may be adjusted independently of each other. 
Engines of this kind must necessarily be supplied with a receiver 
or storage space in which the exhaust steam from the high-pres- 
sure cylinder may be stored if cut-off in the low-pressure cylinder 




Fig. 150. 

does not occur at the same time as compression in the high- 
pressure cylinder. If the cylinders are placed near each other so 
that the connecting passages are short, the receiver is usually 
in the form of a separate vessel connected in the passage between 
the two cylinders; but when the cylinders are some distance 
apart, the passage connecting the two cylinders has enough 
volume to act as a receiver, and no separate vessel is necessary. 
Tandem-Compound Engines. — The presence of a receiver 
modifies somewhat the action of the steam in the cylinders from 
that described above and illustrated in Fig. 149. For a tandem- 
compound engine in which the connecting passage acts as a 
receiver, or for a cross-compound engine with cranks 180° apart 
and supplied with a receiver, the action of the steam in the cylin- 
ders may be shown by the diagrams in Fig. 150, which are similar 



COMPOUND ENGINES 233 

to those shown in Fig. 149 except that cut-off in the low-pressure 
cyHnder does not occur at the same time that compression occurs 
in the high-pressure cyhnder. 

In this case it will be observed from Fig. 150 that cut-off in 
the low-pressure cylinder occurs a little after half stroke and 
considerably before compression (marked D) occurs in the high- 
pressure cylinder. When cut-off occurs in the low-pressure 
cylinder, the steam then in that cylinder expands in the usual 
manner. The high-pressure cylinder, however, has not finished 
exhausting at this time; hence the remainder of the exhaust 
from the high-pressure cylinder is stored in the receiver. Since 
no steam is being drawn from the receiver at this time, the 
pressure in it, which is also the exhaust pressure of the high- 
pressure cylinder, increases as shown by the line CD in Fig. 
150. At D compression occurs in the high-pressure cylinder and 
the exhaust valve closes communication with the receiver. 

The point of cut-off in the low-pressure cylinder controls the 
increase of pressure in the receiver, from C to D, the increase of 
pressure being greater with an early cut-off and smaller with a 
later cut-off. The cut-off in the low-pressure cylinder must be 
so timed that the pressure in the receiver will be the same at D 
as at Ej the point where the exhaust valve on the high-pressure 
cylinder opens. If the pressure at D is not so high as at E, 
the pressure at the end of expansion in the high-pressure cylinder 
will be greater than that in the receiver and there will be a drop 
of pressure the next time the exhaust valve on the high-pressure 
cylinder opens. This would cause a waste of pressure and a loss 
of work, which is to be avoided if possible. 

Cross-Compound with Receiver. — The action of steam in the 
cylinders of a cross-compound engine with cranks set 90° apart 
presents another interesting case. An engine of this kind must 
necessarily be supplied with a receiver because one piston is at 
mid-stroke when the other is at the end of its stroke; hence, 
exhaust from the high-pressure cylinder progresses for one-half 
of a stroke when no steam is being admitted to the low-pressure 
cylinder, and it is necessary to have a receiver in which to store 
this steam. 

The diagrams from the high- and low-pressure cylinders of an 
engine of this type are shown in Fig. 151. These diagrams are 
not drawn in the usual manner, but instead, the low-pressure 
diagram is displaced one-half stroke from the high-pressure 



234 



STEAM ENGINES 



diagram in order to show the relative pressures in the cyHnders 
at any instant. 

It will be observed from Fig. 151 that the exhaust pressure in 
the high-pressure cylinder increases gradually from the beginning 
to the middle of the exhaust stroke. The reason for this is that 
during this part of the stroke the high-pressure cylinder is ex- 
hausting into the receiver and the low-pressure cylinder is not 
taking any steam from it; hence the exhaust pressure in the 
high-pressure cylinder, which is also the receiver pressure, in- 
creases. When the high-pressure piston reaches mid-stroke, 
the low-pressure cylinder begins to admit steam, since the cranks 
are 90° apart, and the receiver pressure is reduced. Thus, the 




Fig. 151. 

high-pressure exhaust line rises from beginning to mid-stroke 
and falls from mid-stroke to the point of compression. 

The admission line for the low-pressure cylinder follows the 
shape of the last half of the exhaust line of the high-pressure 
cylinder; hence it shows a decreasing pressure. In the low- 
pressure diagram shown in Fig. 151 cut-off occurs at or before 
mid-stroke, or before the high-pressure piston has completed 
its stroke. If cut-off in the low-pressure cylinder occurs after 
half stroke the high-pressure piston will have started on its return 
stroke and exhaust will have commenced from the other end of 
the cylinder; hence the pressure in the receiver will again begin 
to increase and this will produce a corresponding increase in the 
admission pressure for the low-pressure cylinder. The effect 
of the second admission of steam into the receiver before the low- 



COMPOUND ENGINES 235 

pressure cut-off is illustrated in Fig. 152, where the admission 
pressure for the low-pressure cylinder is shown decreasing up to 
mid-stroke and increasing from mid-stroke to the point of cut- 
off. This second increase in pressure is called ''second admis- 
sion," and is to be found only when cut-off in the low-pressure 
cylinder occurs after mid-stroke. 

One of the advantages of the cross-compound engine with 
cranks 90° apart is illustrated by Fig. 151 which shows that the 
range of temperature in it is less than in the cylinders of a cross- 
compound with cranks set 180° apart (Fig. 149), or a tandem- 
compound (Fig. 150), because the exhaust from the high-pressure 
cylinder shows a more uniform pressure. The variations in 
pressure illustrated in Figs. 149, 150, 151, and 152 will not show 
to such a marked degree on the actual indicator diagrams because 



Fig. 152. 

the high-pressure diagram is drawn with a stiff indicator spring. 
The variations in pressure in the low-pressure admission may be 
detected easily on the actual diagram, however, because this 
diagram is drawn with a weak spring. 

Power of a Compound Engine. — The power developed by any 
steam engine, whether simple or compound, depends upon the 
number of times the steam is expanded, that is, upon its ratio 
of expansion. It evidently does not matter, then, as far as the 
power of a compound engine is concerned, whether the total 
expansion of the steam occurs in one cylinder or is divided 
between two cylinders, provided only that the steam is expanded 
the same number of times in each case. 

In a compound engine, the total expansion is divided between 
two cylinders for the purpose of reducing cylinder condensation, 
and not for the purpose of increasing the power of the engine. 
The total power developed in both cylinders of a compound 
engine could be developed in the low-pressure cylinder alone 



236 STEAM ENGINES 

by having cut-off in the low-pressure cyhnder occur early enough 
to secure as many expansions of the steam in this cylinder as was 
secured in both the high- and low-pressure cylinders. For exam- 
ple, if cut-off in the high-pressure cylinder occurs at one-quarter 
stroke the steam will expand approximately four times in this 
cylinder. If the volume of the low-pressure cylinder is three 
times that of the high-pressure cylinder then the total expansion 
of the steam will be 

4 X 3 = 12 

This number of expansions could have been secured in the low- 
pressure cylinder alone by admitting the steam directly to that 
cylinder and having cut-off occur at approximately 3^f 2 stroke. 

In any case the approximate ratio of expansion in a multiple 
expansion engine may be found by multiplying the ratio of expan- 
sion in the high-pressure cylinder by the ratio of the volume of 
the low-pressure to the high-pressure cylinder, or by the ratio 
of the square of their diameters. In order to find the ratio of 
expansion more accurately, the clearance volumes would have 
to be considered, but this is not necessary for ordinary purposes, 
as the ratio of expansion changes with the cut-off which, in turn, 
varies with the load. 

From the above discussion it will be seen that the total power 
developed by a compound engine depends upon the ratio of 
the cylinder volumes and upon the point of cut-off in the high- 
pressure cylinder. For any given engine the ratio of the cylinder 
volumes is a fixed quantity, therefore we may say as a general 
proposition that the power developed by a compound engine 
depends only upon the point of cut-off in the high-pressure cylinder. 

The point of cut-off in the low-pressure cylinder has no effect 
whatever upon the total amount of work done by a compound 
engine. The point of cut-off in the low-pressure cylinder does, 
however, control the distribution of work between the two cylinders. 
It does this by affecting the exhaust pressure of the high-pressure 
cylinder. If cut-off in the low-pressure cylinder occurs early in 
the stroke, the exhaust pressure of the high-pressure cylinder 
will be high and the work performed in this cylinder will be a 
smaller portion of the total work and the work performed in the 
low-pressure cylinder will be a larger portion of the total work. 
On the other hand, if cut-off in the low-pressure cylinder occurs 
late, the exhaust pressure of the high-pressure cylinder will be 



COMPOUND ENGINES 



237 



lower and a larger proportion of the total work will be performed 
in the high-pressure cylinder. 

The low-pressure cut-off should be adjusted so as to secure 
an approximately equal division of work between the high- and 
low-pressure cylinders, and, also, so there will be a small drop in 
pressure at the end of expansion in the high-pressure cylinder 
when the engine is working under load. The object in having 
a small drop of pressure at the end of expansion is that there will 
be no gain in carrying the expansion completely down to exhaust 
pressure and, moreover, a little drop in pressure into the receiver 
is needed to secure a quick flow of steam out of the high-pressure 



[steam: 140^*" Gauge 
Normal Conditions-. < Vacuum: 25" to 27" Hg. 
I Speed: 120 rp.m. 




Scale :80;*^=r 
Cylinder: 14-"x 36" 




Vacuum Line 



HP DIAGRAMS 




Vacuum Line 



L.R DIAGRAMS 
Fig. 152a. 



cylinder. However, when the engine is working under no load 
or only a small load there should be no drop of pressure into the 
receiver, but instead, the receiver pressure should be higher than 
the pressure at the end of expansion in the high-pressure cylinder. 
The importance of this should not be overlooked, because, if the 
receiver pressure becomes too low, a condition may be produced 
under which the engine will run away. 

Such a condition as this is illustrated in Figs. 152a and 1526. 
These illustrations show the indicator diagrams from a cross- 
compound engine in which the condition mentioned above 
existed. Fig. 152a shows the indicator diagrams from the high- 
pressure cylinder at A and those from the low-pressure cylinder 
at A', both being taken while the engine was running under no 



238 



STEAM ENGINES 



load. It will be noted by examining these diagrams that the 
receiver pressure is too low, as indicated by the exhaust Hne of 
the high-pressure diagrams and also by the admission lines of the 
low-pressure diagrams. The exhaust pressure of the high-pres- 
sure cylinder is so low that it is impossible for negative work to be 
done in the high-pressure cylinder, and, even though the governor 
is causing cut-off at the earhest possible point, the expansion 
of steam in the high-pressure cyhnder is doing more work than 
needed to carry the friction load at normal speed. Consequently 
the speed increases. When the speed had reached 140 R.p.m. the 




{Steam-. 14-0 Gauge 
Vacvvm. 25" to 27" Hg 
Speed: 120 r.p.m. 

E> 

Scale. 80*= I " 
Cylinder: 14" x 3©" 




Atmospheric Line 



>phe 
Vacuum Line 



H.P. DIAGRAM 



Atmospheric Line 




B 
Scale: 20 
Cylinder: 28 



Mi^L 1 1 ^^ ^^^^ ^^^^.y^^^ ,^ , ^ 1 1^1^ I n 



Vocuum Line 



.. v>v , v.^^^^^^^^^^'^^^^^^^SSSS^^^^ ^^^^.^\^^^^^^^^^^^ 



LP. DIAGRAM 

Fig. 152&. 

engine was stopped by closing the throttle, but if this had not 
been done the speed would have continued to increase and the 
engine would have run away. If the cut-off in the low pressure 
had occurred earlier in the stroke, the receiver pressure, which 
is the exhaust pressure of the high-pressure cylinder, would have 
been higher and negative work would have been done in the high- 
pressure cylinder. This would have put sufficient load on the 
high-pressure cylinder to prevent the speed from increasing above 
its normal value. 

Fig. 1526 shows diagrams taken from the engine when running 
under no load but with the cut-off in the low-pressure cylinder 
adjusted to occur earlier. It will be observed that the receiver 



COMPOUND ENGINES 239 

pressure is now 7H to 8 lbs. above atmospheric pressure and the 
expansion in the high-pressure cyHnder carries the pressure below 
receiver pressure so that negative work (indicated by the cross- 
hatched lines) is done by the high-pressure piston during its 
exhaust stroke. This negative work is sufficient to hold the 
speed down almost to normal. 

Advantages and Disadvantages. — The principal advantage 
derived from compounding is the reduction of losses resulting 
from cylinder condensation and reevaporation. With a simple 
engine these losses increase with high-steam pressures and with a 
large number of expansions of the steam. Hence, a simple 
engine is not well adapted to the use of high pressures nor for 
an early cut-off, both of which are necessary for the economical 
use of the steam. It will now be understood why a compound 
engine is much better adapted for the use of high-pressure steam 
and for expanding the steam a large number of times, hence the 
general use of this type of engine for producing large amounts of 
power, when efficiency is a very important factor. 

Most of the disadvantages of the compound engine are of a 
mechanical nature and arise from the greater complication of this 
type. The greater number of parts make them more expensive in 
first cost and also make them more expensive to maintain than a 
simple engine, on account of more repairs being necessary. The 
greater number of moving parts also adds to the cost of lubrica- 
tion and increases the loss of power in friction. In these respects 
triple expansion and quadruple expansion engines are at even 
greater disadvantage than compound engines, with the result 
that quadruple expansion engines have dropped out of use for 
stationary purposes and the use of triple expansion engines is 
confined almost entirely to large pumping plants. 



CHAPTER XVIII 
CONDENSING APPARATUS 

Purpose of the Condenser. — When an engine exhausts into the 
atmosphere, the exhaust stroke of the piston is made against the 
atmospheric pressure, which acts upon the entire face of 
the piston. This pressure acts in a direction opposite to that in 
which the piston is moving and tends to retard its motion. The 
piston must overcome not only the atmospheric pressure but 
also the friction of the exhaust steam in passing through the 
ports and exhaust pipe on its way from the cylinder to the 
atmosphere. The atmospheric pressure (14.7 lb. per sq. in.) 
added to the friction of the exhaust passages makes a total pres- 
sure of between 15 and 20 lbs. per sq. in. which the piston must 
move against. When it is considered that this back pressure 
acts upon the piston during almost the entire exhaust stroke, 
and that the piston must do work in moving against this pressure, 
it will be realized that the engine could do considerably more 
useful work if this back pressure were removed. 

Removing or reducing the back pressure on an engine increases 
its mean effective pressure. When it is remembered that the 
mean effective pressure of an engine is directly proportional to its 
indicated horsepower, it will be seen that anything which in- 
creases the mean effective pressure will also increase the indicated 
horsepower developed by the engine. In order to gain some idea 
of the increase in horsepower by lowering the back pressure 
consider an engine taking steam at an absolute pressure of 100 
lb. per sq. in., cutting off at J^ stroke, and exhausting into the 
atmosphere against a back pressure of 16 lb. per sq. in. The 
theoretical mean effective pressure under these conditions will 
be 43.7 lb. per sq. in. If the back pressure was reduced to 1.7 
lb. per sq. in. (26 in. vacuum) the mean effective pressure would 
be increased to 

43.7 + (16 - 1.7) = 58 lb. per sq. in. 

^vhich would result in an increase of power of 

240 



CONDENSING APPARATUS 241 

100 ^^~t^''' = 32.8 per cent. 
43.7 

The actual gain would be somewhat less than this depending 
upon the type of engine and the conditions of operation, but 
in any case it would be considerable. 

The back pressure on an engine is reduced by leading the 
exhaust steam into a closed vessel and condensing it into water, 
instead of permitting the engine to exhaust directly into the 
atmosphere. Such a closed vessel is called a condenser. As the 
exhaust steam enters the condenser, it either meets a spray of 
cold water or comes in contact with tubes through which cold 
water is flowing. In either case, the water extracts heat from the 
exhaust steam and condenses it into water. Since the water 
occupies only about Mtoo of the space occupied by the exhaust 
steam, the pressure in the condenser is reduced by the condensa- 
tion of the steam. In order to maintain the low pressure in 
the condenser it is necessary to condense the exhaust steam as 
fast as it enters and to constantly remove the water and any air 
which may come in with the exhaust steam. 

The purpose of installing a condenser may be either to increase 
the efficiency of the engine or to increase the power of the engine. 
A condensing engine will be more efficient than a noncondensing 
one for the reason that cut-off in the condensing engine may be 
shorter than in the noncondensing engine when the same amount 
of power is developed in both, on account, of the great number of 
times the steam is expanded in the condensing engine. For 
example, an engine running noncondensing may cut off at 3^^ 
stroke and develop a certain amount of power. The same 
engine connected to a condenser may cut off at J^ stroke and 
develop the same amount of power. Since the amount of steam 
used by the engine is in proportion to the cut-off, the engine will 
use M — J^ = /i less steam when running condensing than 
when running noncondensing. The amount of steam which can 
be produced from a pound of coal is ordinarily from 7 to 10 
pounds, but the amount of power obtained from the steam 
depends upon how the steam is utilized. Since, by running an 
engine condensing rather than noncondensing the steam is 
utilized more efficiently, power plant engines are almost invari- 
ably run condensing unless the exhaust steam is used for heating. 

A Corliss engine running noncondensing will use from 25 to 30 
pounds of steam per indicated horsepower per hour but if run 

24 



242 STEAM ENGINES 

condensing, it will use only about 20 pounds. For compound 
engines, the amount used will be about 25 pounds noncondensing 
and about 15 pounds condensing. A triple expansion engine 
running condensing will produce an indicated horsepower from 
as little as 10 pounds of steam. 

While a condensing engine will require from 20 to 30 per cent, 
less steam than a noncondensing one, this apparent decrease in 
steam consumption does not represent a net gain. The steam 
used by the condenser pumps must be added to that consumed 
by the engine unless the exhaust from the pumps is used for 
heating the feed water in which case only the difference between 
the heat entering and leaving the pumps should be charged to 
the engine. 

Condensation of Steam. — The condensation of steam is just 
the reverse of the process by which steam is formed, and the 
amounts of heat involved are the same; the only difference being 
that heat must be added to water to change it into steam and 
that heat must be taken away from the steam to condense it into 
water. Moreover, for the same conditions of pressure, quality, 
and weight of steam exactly the same amount of heat must be 
transferred in either case, being transferred i7ito the steam in one 
case and out of it in the other. 

A pound of steam at any pressure contains a definite amount of 
latent heat of evaporation, as may be seen by reference to the 
steam table in Chapter 5. If this amount of heat is taken out of 
the steam, a pound of it condenses into water and the water will 
have the same temperature as the steam from which it was con- 
densed. If only one-half of the latent heat in a pound of steam 
is extracted, then only one-half of a pound of steam will be con- 
densed and the resulting water will be at the same temperature 
as the steam. The same is true for any amount of heat taken 
from the steam. The weight of steam condensed will be the 
number of heat units extracted divided by the latent heat of one 
pound of steam at the pressure of condensation. If the exhaust 
steam is wet, that is, contains moisture suspended in it, this 
moisture contains no latent heat, therefore only the latent heat 
actually contained in the steam must be extracted in order to 
condense it. While steam may not be entirely dry at the end 
of expansion, the drop in pressure at release usually completes 
the drying, so that in calculations relating to condensers it is 
usually assumed that the exhaust steam is dry. 



CONDENSING APPARATUS 243 

In order to condense steam it is necessary to bring it into con- 
tact with something which has a lower temperature because heat 
will only pass into a substance at lower temperature. For this 
reason the condensing water used in a condenser must have a 
lower temperature than the exhaust steam that is to be condensed. 
When the condensing water absorbs heat from the exhaust steam, 
the steam is condensed and the temperature of the condensing 
water is increased. It is evident that the steam cannot be con- 
densed unless its temperature is higher than the final temperature 
of the condensing water. 

By continually condensing the steam in the condenser a low 
pressure is maintained in it. The steam is expanded in the 
cylinder almost to this pressure, and when the exhaust valve 
opens, the steam pressure drops to the same pressure as that in 
the condenser. At the same time, its temperature drops to that 
shown by the steam table to correspond to its pressure. For 
example, suppose the absolute pressure in the condenser is 
maintained at 2 lb. per sq. in. then the exhaust steam entering 
it will have this pressure and it will have a temperature of 126.15° 
F. as will be seen by referring to the steam table in Chapter 5. 
The exhaust steam at this pressure and temperature has a latent 
heat of 1021 B.t.u. per pound, and in condensing it gives up this 
heat to the condensing water. The condensate, or water result- 
ing from the condensation of the steam, will also have a tempera- 
ture of 126.15°F. 

Measuring Vacuum. — Strictly speaking, a vacuum means a 
space in which there is no pressure, or in which the absolute 
pressure is zero. However, in steam engineering work the word 
vacuum refers to any space in which the pressure is less than 
atmospheric pressure. For this reason, the reduced pressure in a 
condenser is called a vacuum. 

The vacuum in a condenser may be measured by means of a 
mercury column or by means of a gage constructed somewhat 
like a pressure gage but marked to read pressures less than that 
of the atmosphere. The mercury column is the more accurate 
method and is generally used where the pressure in the condenser 
is very low. 

A device for measuring vacuum by means of a mercury column 
is illustrated in Fig. 153. In this device a glass tube about 80 
inches long is bent into a U-shape, and is filled about half full of 
mercury. One branch of the glass U-tube is connected to the 



244 



STEAM ENGINES 



space in which the vacuum is to be measured, the other branch 
being left open so that it is under atmospheric pressure. As the 
pressure is reduced in the space into which the U-tube is connect- 
ed (in this case, a condenser), the mercury will rise in that branch 
to a height A, corresponding to the difference in pressure on the 
surfaces of the mercury in the two branches of the U-tube. 

The amount of the vacuum is usually expressed in inches of 
mercury, or simply ''inches/' and is the difference in height of 
mercury in the two branches of the U-tube. Thus, if the height 
A in Fig. 153 is 20 inches, the vacuum amounts to 20 inches of 
mercury, or is said to be ''20 inches." It should be remembered 
that the height of the mercury column indicates the reduction of 
pressure, and not the actual pressure existing in the condenser. A 




Fig. 153. 

vacuum of 20 inches means that the pressure has been reduced 
enough to support a column of mercury 20 inches high. Since 
a column of mercury 1 inch high is equivalent to a pressure of 
.49 lb. per sq. in., 20 inches corresponds to a reduction of pressure 
of 20 X .49 = 9.8 lb. per sq. in. below atmospheric pressure. 
Before the pressure still existing in the condenser can be found, 
it is necessary to know the pressure of the atmosphere. If the 
atmospheric pressure is 14.7 lb. per sq. in., a vacuum of 20 inches 
leaves a pressure of 14.7 — 9.8 = 4.9 lb. per sq. in. If the 
barometer, which measures the atmospheric pressure, reads 
28 inches, then 20 inches of vacuum leaves a pressure of 28 — 20 
= 8 inches of mxercury, or 8 X .49 = 3.92 lb. per sq. in. 

It is seen from the above discussion that a statement to the 
effect that the vacuum carried by a condenser is a certain number 
of inches does not always mean the same thing, because of varia- 



CONDENSING APPARATUS 245 

tion in the atmospheric pressure. Thus, a vacuum of 22.5 inches 
at a place 5280 feet above sea level is as near a perfect vacuum as 
28 inches at New York, which is at sea level. In the first men- 
tioned place 24.5 inches would be a perfect vacuum, while at 
sea level 30 inches would be a perfect vacuum. Vacuum gages 
of all kinds are practically always marked to read in inches of 
mercury to correspond with the U-tube described above. 

In the operation of a condenser there will always be some 
pressure in the condenser due to the presence of water vapor and 
air, both of which exert a pressure. The amount of this pressure 
will depend upon the temperature of the condensate and the 
temperature and quality of the air present in the exhaust steam. 
Air enters the boiler in the feed water and passes into the piping 
system with the steam. It also leaks into the low-pressure 
cylinder of the engine around the piston rod. Leaks in the 
condenser itself and exhaust piping, both of which are under a 
pressure less than that of the atmosphere, also account for the 
presence of some air in the condenser. Whatever air is present 
adds its pressure to that of the water vapor, so that the total 
pressure in the condenser is equal to the sum of the vapor pressure 
and the air pressure. 

The vapor pressure depends upon the temperature of the 
condensate and is the same pressure as that shown in the steam 
table as corresponding to the various temperatures. For exam- 
ple, if the temperature of the condensate is 101.83°F., the vapor 
pressure in the condenser is 1 lb. per sq. in. which corresponds to 

-^Q or 2.04 inches of mercury or to a vacuum of 30 — 2.04 = 

27.96 in. (atmospheric pressure being 14.7 lb. per sq. in.). If 
the temperature of the condensate was 126.15°F., the vapor 
pressure would be 2 lb. per sq. in. corresponding to a vacuum of 

2 

30 Vq or 25.92 in. It must be remembered that the pressure 

of the air in the condenser acts in addition to the vapor pressure, 
so that if, for example, there were enough air present in the con- 
denser, in the last case mentioned above, to create a pressure of 
1 lb. per sq. in. the total pressure in the condenser would be 3 

3 

lb. instead of 2 lb. and the vacuum would be 30 — ^ or 23.87 

in. 



246 



STEAM ENGINES 



It is necessary continually to remove the air and condensate 
as they would soon accumulate and destroy the vacuum. But 
no matter how perfect the condensing and pumping apparatus 
there will always be some vapor pressure due to the temperature 
of the condensate, and there will always be more or less air enter- 
ing the condenser with the exhaust steam. The more perfect 




j&m 



w ufx^mw 



^^^^^^^^^^^^^^^^'^^^^^^^^^^^^'.^^^■^ 




Fig. 154. 



the apparatus for removing these, and the colder the condensing 
water, the higher will be the degree of vacuum that may be 
carried. 

Forms of Condensing Apparatus. — Condensers may be divided 
into two different classes or types; namely, those in which the 
steam is condensed in direct contact with water, the condensate 



CONDENSING APPARATUS 247 

and condensing water mixing and leaving the condenser at the 
same temperature; and those in which the steam is condensed in 
contact with tubes through which or around which the condensing 
water flows. In the latter class the condensate and condensing 
water are kept separate and the condensing water leaves the 
condenser at a somewhat lower temperature than the condensate. 
Practically all condensers fall into one or the other of these two • 
classes although there are many varieties in each class. 

Jet Condenser. — One of the simplest forms of condensers is the 
jet condenser illustrated in Fig. 154. In this condenser, the 
condensing water enters the injection pipe at B. At the end of 
the injection pipe is an adjustable cone-shaped spray head which 
breaks the water into a fine spray. The exhaust steam entering 
at A meets the spray of water, condenses, and falls to the bottom 
of the condensing chamber F. From the bottom of the con- 
densing chamber the mixture of condensing water, condensate, and 
air is taken into the steam-driven wet air pump G and forced out 
through J into the hot well. If there should be an excess of con- 
densing water, due either to the improper regulation of the spray 
head or to the failure of the pump to remove the water as fast 
as it collects in the bottom of the condensing chamber, the water 
will rise in the condensing chamber until it covers the spray head 
and the condensation of steam will be practically stopped. The 
vacuum will then be greatly reduced, or broken, and the pressure 
of the exhaust steam acting upon the surface of the water will 
force it through the valves of the pump and into the hot well. 
There would then be no danger of water flooding back into the 
engine cylinder and wrecking it. 

With this type of condenser it is not necessary to have the 
supply of condensing water under pressure as the vacuum in the 
condensing chamber will draw the condensing water through the 
injection pipe unless the supply is more than about 15 feet below 
the condenser. However, if the condenser draws in its own supply 
of condensing water there must be some means provided for 
creating a vacuum to start the condenser. This may be done 
by starting the pump or by providing an auxiliary injection of 
water under pressure so that the first steam to come from the 
engine may be condensed and thus create a vacuum for starting 
the water through the spray head. 

The amount of vacuum that may be maintained in a jet 
condenser depends upon the temperature of the water in the 



248 



STEAM ENGINES 



bottom of the condenser, the amount of air carried into the con- 
denser by the condensing water and steam and upon the tightness 
of the valves and joints. 

Siphon Condensers. — A type of jet condenser known as the 
siphon condenser is illustrated in Fig. 155. In this condenser 
the cooHng water enters at the side through the pipe A arid over- 
flows the edges of the cone- 
shaped nozzle H. This causes 
the water to form a hollow cone 
which becomes a solid stream in 
passing through the throat of the 
nozzle at E. The amount of 
water passing through the con- 
denser is regulated by means of 
the valve D. The exhaust steam 
enters the condenser at B and is 
given a downward direction by 
means of the goose neck C. It 
then comes in contact with the 
hollow cone of water and is con- 
densed. The mixture of condensed 
steam and condensing water, to- 
gether with any entrained air, 
passes through the contracted 
neck of the cone at E with a 
high velocity and is discharged 
into the tail pipe at F, the lower 
end of which is below the surface 
of the water in the hot well. 

When in operation, with a 
vacuum in the condenser, the tail 
pipe filled with water, and the end of the tail pipe sealed by the 
water in the hot well, which is under atmospheric pressure, the 
condenser acts like a barometer. The atmospheric pressure on 
the surface of the water in the hot well would therefore force the 
water in the tail pipe to a height of 34 feet which corresponds to 
atmospheric pressure. For this reason, it is necessary that the 
tail pipe be at least 34 feet long in order to prevent the water 
from backing through the condenser and into the cylinder of 
the engine. The force of the water passing through the con- 
tracted neck of the nozzle will balance several pounds of atmos- 




FiG. 155. 



CONDENSING APPARATUS 



249 



pheric pressure so that tlie length of the tail pipe might be made 
a little less than 34 feet but, to guard against any possibility of 
an accident, the tail pipe is made at least 34 feet long. 

The condensing water may be pumped into the condenser under 
pressure or the vacuum in the condenser may draw it in if it does 
not have to lift it more than about 15 feet. It is possible to use 
very muddy or dirty water in this type of condenser as it is not 



Hi 



yf/fjf//ffni >>/i/> i-m' 



FROM ENOINE. 



^)niini>, 31,13, ,),,, ,,. 




Fig. 156. 

very likely to become stopped up. However, should a stoppage 
occur or should the condenser fail to operate for any reason, 
pressure will accumulate in the condenser until it opens the auto- 
matic relief valve shown at G in Fig, 155, and the engine will 
then exhaust into the atmosphere. 

Barometric Condenser.- — Although the siphon condenser just 
described might be classed as a barometric condenser inasmuch 
as there is barometric action in the tail pipe, the name barometric 
condenser is usually reserved for that class of condensers in which 
there is no injector or ejector action through a contracted tube. 

25 



250 STEAM ENGINES 

Such a condenser as this is illustrated in Fig. 156, the tail pipe 
being left off to permit a larger illustration being used. 

In this type of condenser the exhaust steam is brought into 
contact with a spray of cold water and is condensed. The mixture 
of condensed steam and water then flows by gravity through a 
tail pipe, the lower end of which is sealed by the water in the hot 
well. In operation, the tail pipe of this condenser acts as a true 
barometric tube with the water standing in it at a height corre- 
sponding to the vacuum carried in the condenser. As more water 
is added to the column at the top, a like amount flows out at the 
bottom. 

The condensing water enters the injection pipe at B and is 
broken into a spray by means of the cone F. This cone is sus- 
pended from a coil spring above so that the nozzle opening is 
automatically adjusted by the water pressure. The exhaust 
steam enters at A and divides into two parts, one part passing 
directly into the condensing chamber D where it meets the spray 
of cold water and the other part passing downward through the 
annular space E and then upward into the condensing chamber. 
Any air that is not entrained in the water passing into the tail 
pipe and also any uncondensed vapor pass up through the center 
of the spray nozzle, which is hollow, and into the air cooling 
chamber at the top. The air is cooled and the vapor condensed 
by an injection of cold water through the pipe K. The air is 
then drawn off through the air pipe H by means of a dry air 
pump. Since the air and water move in opposite directions, this 
type of condenser is known as a countercurrent condenser. 

Surface Condensers.^ — This class of condensers includes all 
those in which the steam and condensing water are kept separate 
and in which the steam is condensed on a metal surface. These 
condensers usually consist of a shell which contains a large number 
of small tubes, with the condensing water on the inside of the 
tubes and the steam on the outside, although in a number of 
makes of surface condensers the condensing water is on the 
outside and the steam on the inside of the tubes. The tubes are 
usually made of brass because this metal conducts heat better 
than iron. 

A typical example of a surface condenser is illustrated in Fig. 
157 which shows a Wheeler condenser with wet air pump and 
circulating pump for forcing the condensing water through the 
condenser. This condenser, which has a rectangular cross section, 



CONDENSING APPARATUS 



251 









o 



252 STEAM ENGINES 

consists of a cast-iron shell or case containing a large number of 
closely spaced small seamless drawn brass tubes through which 
the condensing water circulates. The tubes have considerable 
length compared with their diameter and they are therefore 
supported at their middle points to prevent their sagging. They 
are fastened into the tube sheets at their ends by means of a 
stuffing box made of a brass ferrule with packing. This con- 
struction is used so that the tubes will be free to expand and 
contract without leaking and so they may be easily removed. 

The space in the shell between the tube sheets and the heads 
of the condenser forms the condensing water compartments. 
The water compartment at the right of Fig. 157 is divided by a 
baffle plate which causes the water to flow in one direction through 
the bottom set of tubes and to flow in the opposite direction 
through the top set, the water being forced through the tubes 
by the pump shown at the right of Fig. 157. 

The exhaust steam enters the condenser through the large 
opening shown at the top of the shell. Upon entering, it strikes 
the baffle plate shown just over the tubes, which prevents it from 
striking directly on the tubes and distributes it over a greater 
area of cooling surface. In this way the entire tube surface is 
made more effective. The condensation of steam begins on the 
surface of the upper row of tubes and continues as the steam 
passes down among the tubes, the condensate dropping to the 
bottom of the condenser shell. The condensing water entering 
at the bottom, increases in temperature as it flows towards the 
top, until its temperature upon leaving is practically the same as 
that of the exhaust steam. The condensate, together with any 
entrained air, is drawn into the wet vacuum pump shown at the 
left of Fig. 157 and is pumped out of the condenser. 

In operating a condenser of the kind described above particular 
care should be used to see that exhaust steam is not turned into 
the condenser unless there is water in the tubes as the tube pack- 
ings are liable to be destroyed and the shell or tube sheets cracked, 
particularly if cold water is admitted. Sudden or large changes 
of temperature should be avoided as they are apt to injure the 
tubes or to cause them to leak. The water chambers should be 
examined frequently to see that no foreign matter is collecting 
and stopping the tubes as this will lower the vacuum and increase 
the duty of the tubes and circulating pump. The outside of the 
tubes can be kept clean and efficient by introducing about a 



CONDENSING APPARATUS 253 

gallon of kerosene with the exhaust steam once a week and just 
before shutting down. This will free the tubes from oil carried 
over in the exhaust steam. 

High Vacuum Condensers. — Any of the condensing systems 
described above are well adapted for use with steam engines be- 
cause a steam engine operates at its best commercial economy 
with a vacuum of about 26 in. (referred to a 30-in. barometer) 
and any of these condensers will produce a vacuum of from 24 to 
26 in. If a steam engine is operated at a higher vacuum than 
about 26 in., cylinder condensation becomes excessive because of 
the large size of the cylinder made necessary by the largely in- 
creased volume of the steam at very low pressures. The in- 
creased size of the piston also increases the friction so that the 
loss from these two sources more than balances the gain from 
increased expansion of the steam. 

A steam turbine, on the other hand, will show a considerable 
increase in efficiency from the use of vacua above 26 in. This is 
largely because the steam in passing through the turbine experi- 
ences a gradual drop in temperature and the parts of the turbine 
are not subjected to a great range of temperature as in a steam 
engine, hence there is not much condensation of steam in the 
turbine. For the reasons just mentioned condenser manufac- 
turers have made considerable improvement in condensing sys- 
tems since the introduction of the steam turbine. The principal 
improvements in securing higher vacua were the use of a separate 
dry vacuum pump for removing the air and noncondensable 
vapors, and also the providing of means for cooling the air before 
it enters the air pump so that the volume to be handled will not 
be so large. 

Choice of a Condenser.^ — The choice of a condenser for a steam 
engine depends largely upon the kind and cost of the available 
supply of condensing water. Where there is a plentiful and 
cheap supply of good condensing water which is also suitable for 
feeding into the boiler, some good type of jet condenser will gen- 
erally be found most desirable. If there is sufficient overhead 
room, a siphon or barometric condenser will be found most desir- 
able and least expensive. In this connection, it must be remem- 
bered that these condensers may often be located to advantage 
outside the building. A condenser of the siphon or ejector type 
will also use to advantage a very dirty or muddy condensing water 
which would not be at all suitable for feeding into the boiler. If 

26 



254 STEAM ENGINES 

this is done, however, the feed water must be heated in some other 
way, but this can usually be done by means of the exhaust from 
the steam-driven pumps and other auxiliaries around the power 
plants. 

When the supply of condensing water is not suitable for feeding 
into the boiler, a surface condenser should be used because then 
the same feed water may be used over and over and only enough 
additional feed water need be supplied to make up the losses by 
leakage and other sources of waste. However, the condensing 
water used in a surface condenser should not be so dirty as to 
cause stoppage of the tubes nor should it contain enough mineral 
matter to give trouble from incrustation as the tubes of a surface 
condenser are so small and so numerous that cleaning them is 
difficult and expensive. On the other hand, not much trouble 
from incrustation is to be expected because the condensing water 
does not reach a high enough temperature to cause mineral sub- 
stances to deposit very fast and also because the velocity of the 
water through the tubes is high enough to retard the deposit of 
mineral matter. 



CHAPTER XIX 
LUBRICATION 

Friction. — Every bearing in every piece of machinery in a 
power plant produces friction. Friction in bearings generates 
heat, which is one form of energy, and, as there is no way of utiHz- 
ing this heat, it is lost. For this reason, friction in bearings 
represents a direct loss. The continual overcoming of friction 
requires power, and it is for this reason that some power is 
required to run a steam engine even when it is carrying no load 
except the load caused by the friction. 

The power lost in overcoming friction in a power plant depends 
upon the number, kind, and condition of the bearings. It is 
seldom less than 5 per cent, of the total power generated and it is 
sometimes as much as 30 per cent. The power wasted by friction 
in a steam engine will amount to from 5 to 20 per cent, of the total 
power of the engine, with about 10 per cent, representing the 
average for an engine with good bearings well lubricated. 

The losses due to friction are not only the loss of power but they 
include also the repairs and depreciation due to wear on the bear- 
ings, guides, piston rod and packing, piston, and other rubbing 
surfaces. 

The losses mentioned above may be reduced considerably by 
using a sufficient quantity of the proper kind of lubricant. The 
selection of the kind of lubricant is, therefore, a very important 
problem and a change in the kind of lubricant used may often 
result in the large increase in the economy of operation of the en- 
gine. The choosing of the proper lubricant to use in any particu- 
lar bearing is not a very simple matter, as there is such a large 
variety of lubricants on the market. The selection of a lubricant 
for any particular purpose should, therefore, be undertaken only 
after the fundamental principles of lubrication and the necessary 
qualities of the lubricant are thoroughly understood. 

Lubrication. — If you should look through a microscope at the 
surface of a polished shaft it would appear to be rough, even 
though as far as the naked eye can isee, the shaft is perfectly 
27 255 



256 STEAM ENGINES 

smooth. After looking at such a surface through a microscope 
and seeing its roughness, it will be realized that when two such 
surfaces are brought together the elevations and depressions of 
one surface will interlock with those of the other surface, as 
shown magnified in Fig. 158, and that considerable force will be 
required to move one surface over the other. In other words, 




Fig. 158. 

there will be considerable friction between the two surfaces. The 
friction in this case would arise from two sources : first, from mov- 
ing the irregularities on the surface of the journal into and out of 
the irregularities on the bearing surface, that is, the friction due 
to the unevenness of the metal surfaces; second, the friction due 
to the cutting action of the metal surfaces upon one another. 




Fig. 159. 

Even though the irregularities in the metal surfaces are invis- 
ible to the naked eye, it will be realized that the amount of fric- 
tion between the surfaces is enormous, especially if the bearing 
supports considerable weight or runs at high speed. 

Principles of Lubrication.^ — The object of lubrication is to place 
a thin film of oil between the metal surfaces so they will not come 



LUBRICATION 



257 



in contact with each other while running. If this object is accom- 
phshed the metal surfaces will not touch but the journal will 
'Afloat on oil" and the friction will be enormously reduced. This 
condition is illustrated in Fig. 159 which shows the two surfaces 
of Fig. 158 but with a film of oil between them. 

In order to secure a film of oil between bearing surfaces some 
provision must be made so that the surfaces will not scrape off the 
film of oil but rather that the moving surface will be made to ride 
upon the particles of oil. When the surfaces to be lubricated are 




Fig. 160. 

flat, as is the case with the crosshead or piston of an engine, 
the ends of the crosshead or piston should be slightly beveled so 
that there are no sharp edges. If the edges are left sharp the film 
of oil will be scraped off and there will be metallic contact between 
the bearing surfaces. This is illustrated in Fig. 160 which shows 
the end of a piston slightly beveled so that oil may collect under 
it and keep the surfaces apart. For the same reason, the edges of 
piston rings should be slightly beveled instead of left with sharp 
edges. 

With a round shaft which turns in a bearing, such as the main 



258 



STEAM ENGINES 



bearing of an engine, the same result is automatically secured 
through the fact that the bearing which surrounds the shaft 
always has a slightly larger diameter than the shaft. This leaves 
a small clearance at the top and sides of the bearing, as illus- 
trated in Fig. 161, which is somewhat exaggerated in order to 
show the clearance more plainly. The clearance at the sides of 
the bearing permits the oil to be drawn in between the journal 
and bearing and thus form a film between them. The oil adher- 
ing to the surface of the journal not only causes it to be drawn 
into the bearing but also causes it to be drawn out at the other 
side. With large bearings which are supplied with oil at the 




Fig. 161. 



center of the top it is necessary to provide diagonal oil grooves in 
the bearing surface so that the oil may be spread to all parts of it. 
There will always be some friction, even with a well-lubricated 
bearing in which a film of oil is maintained, but in this case the 
friction is largely a fluid friction instead of the friction of metal- 
lic surfaces in contact. When the oil has been carried into the 
bearing it comes in contact with the bearing surface and adheres 
or sticks to it. Since the bearing is stationary, the particles of 
oil next to its surface will also be stationary and the particles of 
oil which are next to the revolving journal will be in motion due 
to these particles clinging to the journal. The velocity of the 
particles of oil in the film, therefore, varies all the way from no 



LUBRICATION 259 

velocity on one side of the film to the velocity of the journal on the 
other side of the film. It is this relative motion of the particles 
of oil among themselves that is largely the cause of friction in a 
well-lubricated bearing, and this friction is more of a fluid friction 
than a friction between metal surfaces. 

Characteristics of Oil. — A journal picks up a film of oil and 
carries it into the bearing by reason of the adhesiveness of the 
oil, that is, its ability to adhere to the surface of the journal. The 
quantity of oil that will be drawn into the bearing in this way 
depends upon the viscosity of the oil. A homely illustration 
of this action is the following : If you stick your finger into a cup 
of thick molasses and then withdraw it, a large quantity of the 
molasses will adhere to your finger and be withdrawn with it. If 
you try the same with a cup of water, only a very small quantity of 
water will adhere to your finger and be withdrawn with it, be- 
cause the water is so much thinner and more fluid than the molas- 
ses. When an oil is thick or does not flow readily it is said to be 
viscous, or its viscosity is high. If an oil is thin or flows readily 
its viscosity is low. In other words, the viscosity of oil refers 
to its ''body." 

Since the work which a lubricating oil is called upon to do is 
keeping the metal surfaces of the bearing apart by means of a 
film of oil, the conditions under which the oil is to be used deter- 
mine the proper viscosity or body which the oil should have. 
These conditions are the speed of the moving parts, the weight on 
the bearing and its running temperature. 

If an oil is too low in viscosity it will permit metallic contact 
between the surfaces to be lubricated, with a consequent loss of 
power, and wear on the bearing. On the other hand, if an oil too 
high in viscosity is used, there will be unnecessary fluid friction, 
with a consequent loss of power. The more viscous the oil, the 
greater the pressure which can be sustained without metallic con- 
tact. The viscosity of the oil should, therefore, be in proportion 
to the pressure on the bearing. Heavy, slow-moving bearings 
require an oil of high viscosity; light, swift-moving bearings re- 
quire a thin oil, or an oil of low viscosity. 

The viscosity of an oil is different at different temperatures; 
therefore, in selecting an oil to be used, the temperature at which 
the bearing runs must be considered. Most of the friction in a 
well-lubricated bearing is internal or fluid friction, as explained 
before, and this friction causes the temperature of the bearing to 



260 STEAM ENGINES 

be somewhat higher than the room temperature. Since a heavy 
oil has more internal friction than a light one, the viscosity should 
be only high enough to maintain the film of oil at the temperature 
at which the bearing runs. 

Testing Oils. — From the above discussion of lubrication it is 
apparent that one of the most important qualities of a lubricating 
oil is its viscosity. The measurement of viscosity of lubricating 
oils is in a certain sense unsatisfactory because the results ob- 
tained with the different instruments which are available for this 
purpose do not agree among themselves. For this reason the 
make of instrument used in making the test should be stated in 
quoting the viscosity. 

One of the most common instruments used in testing viscosity 
is the Saybolt viscosimeter. It consists of a tall pipette of 
small diameter surrounded with a jacket which may be used for 
maintaining the oil at any desired temperature during the test. 
The test is made by filling the pipette to a certain point with the 
oil whose viscosity is to be measured and noting the time in sec- 
onds which the oil takes to run out of the pipette. If it is desired 
to find the specific viscosity, that is, the viscosity of the oil com- 
pared with that of water, this may be done by dividing the time 
which the oil takes to run out by the time it takes an equal 
volume of water to run out. It should be particularly noted 
that the viscosity of oil varies greatly with its temperature. 
The temperature at which the test is made should, therefore, be 
stated and, to be of practical use, the temperature during the test 
should be as near as possible to the temperature at which the oil 
is to be used. 

Gumming Test.- — A gumming test is of importance because it 
indicates the extent to which the oil has been refined and also the 
extent to which the oil may be expected to change, due to oxida- 
tion, when in use. 

The gumming test may be made by putting a small quantity of 
the oil to be tested in a small glass vessel, such as a cordial glass, 
and then mixing with it an equal quantity of nitrosulphuric acid. 
A properly refined oil will show little if any change, but a poorly 
refined oil will show the separation of large quantities of material 
of dark color, due to the oxidation of tarry matter contained in the 
lubricant. Oils which contain a large percentage of tar absorb 
the most oxygen and are, therefore, mildly drying. 

Flash and Fire Tests.^ — These tests have nothing to do with the 



LUBRICATION 261 

lubricating qualities of an oil but they give an indication of its 
safety and for this reason they are important. 

The flash test is made by heating the oil slowly in a vessel sur- 
rounded by a proper bath and noting the temperature at which 
a flame passed over the surface of the oil will ignite the vapors 
arising from the oil. The fire test is made by continuing to heat 
the oil slowly and noting the temperature at which the vapors, 
when ignited, will continue to burn. 

Acid Test. — During the process of refining, oil is agitated with 
sulphuric acid for the purpose of removing the tarry matter, and 
this acid must be practically all removed before the oil is put on 
the market as any free acid in it is apt to corrode the bearings in 
which it is used. 

Acid may be readily detected by thoroughly mixing a small 
quantity of the oil with an equal amount of pure warm water. 
The mixture should then be tested with neutral litmus paper. 
If acid is present, the paper will immediately turn red. The 
sensitiveness of the paper may be greatly increased by introduc- 
ing it into the fumes of nitric or hydrochloric acid until it becomes 
partly red. After it is dried and part of its surface is immersed 
in the oil to be tested the very smallest amount of acid will turn 
the part of the paper immersed a shade darker than the rest. 

It should be appreciated by the practical man that the tests 
of lubricating oils give only an approximate idea of the qualities 
of the oil. In fact, no rigid directions can be given for the choice 
of an oil for a specific purpose. It is best to try various lubri- 
cants which can be purchased for any given lubricating problem 
until one is found which gives satisfactory results under actual 
working conditions. This should then be completely tested and 
the results used in buying oil for the same purpose in the future. 

Steam-Engine Lubrication. — The lubrication of a steam engine 
presents two distinct and separate problems. First, there is the 
problem of lubricating such bearings as the main bearings, crank 
and wrist pins, and other external bearings. The principles 
governing the lubrication of these bearings have been discussed 
in the first part of this chapter. Second, there is the problem of 
lubricating the internal bearings, such as piston, valves, piston 
rod, and valve rod. 

The problem of lubricating the internal bearings, especially the 
piston and valves, is far more difficult than that of lubricating the 
external bearings. The external bearings are more easily in- 



262 STEAM ENGINES 

spected to determine if they are being properly lubricated. If 
they receive too little oil the temperature of the bearing becomes 
too high; if they receive too much oil this can be told by the over- 
flow and waste of oil from the bearing. The bearing surfaces of 
the piston and valves are not so easily inspected, nor is the fact 
easily detected that too little or too much oil is being supplied or 
that the oil is being properly distributed. In general, a light oil 
or oil of low viscosity is used in the external bearings, while the in- 
ternal bearings require a heavy oil or oil of high viscosity. Cyl- 
inder oil should also have a high flash and burning point in order 
to prevent it from being carbonized by the high temperature in 
the cylinder. This is especially true when the engine is supplied 
with superheated steam. 

The best method of carrying oil into a steam-engine cylinder 
for internal lubrication is to introduce it into the steam line sup- 
plying the engine. This is accomplished by connecting the feed 
pipe from the lubricator to the steam line, allowing the feed pipe 
to extend into the center of the steam line. By doing this, the 
cylinder oil, coming in contact with the column of steam flowing 
through the steam line at its point of greatest velocity, is broken 
up into very small particles and carried along with the steam into 
the cylinder. Since the steam comes in contact with all internal 
surfaces requiring lubrication, this method insures a supply of 
oil to all rubbing surfaces. The point at which the feed pipe 
enters the steam line should be on the boiler side of the main stop 
or throttle valve so the spindle of the valve will be lubricated, 
thus making easy the operation of the valve. However, the feed 
pipe should not enter at such a point that the spraying or atomi- 
zation of the oil will be affected by a steam separator or angles in 
the pipe between the point of introduction and the steam chest, 
and the oil should be introduced at a point not further than about 
twelve feet back of the throttle valve. 

Superheated steam, which is very dry and hot, is not a good 
carrying medium for cylinder oil, and for this reason the oil should 
be introduced just back of the throttle valve when superheated 
steam is used. 

Steam is sometimes used for jacketing the cylinder and heads 
of the engine before entering the cylinder. Under these condi- 
tions some of the oil will be deposited on the walls of the passages 
if it has been introduced into the steam line in front of the throttle 
valve. To overcome this difficulty oil is introduced directly into 



LUBRICATION 



263 



the valve chambers when the current of steam passing through 
the valves will break up the oil into fine particles and distribute 
it. 

Lubricators. — Cylinder oil is forced into the steam line by 
means of a lubricator. These are of two kinds: hydrostatic 
lubricators and mechanically operated lubricators. 

A hydrostatic lubricator is illustrated in Fig. 162, being shown 
partly cut away so that its internal construction and operation 
may be understood. The body of the lubricator contains cylin- 




FiG. 162. 

der oil and water, the oil remaining on top of the water because it 
is lighter. Water is introduced into the lubricator from the 
condensation of steam in the pipe 3 which is connected to the 
steam line and the bulb 4. It then passes into the bottom of the 
lubricator through the tube 6, where it exerts an upward pressure 
on the oil. This pressure forces the oil out of the lubricator 
through the tube 8 which ends near the top of the lubricator 
where it will be supplied with oil as long as there is any in the 
lubricator. The oil passes through the tube 8 to the regulating 
valve 9 where its rate of flow to the cylinder is controlled. It 
forms into drops as it passes through the regulating valve and 



264 STEAM ENGINES 

these drops rise through the water in the sight feed glass 10 and 
through the pipe 11 into the steam line. As the oil is fed into the 
steam line, its place is taken by more water which enters through 
tube 6; thus as the amount of oil in the lubricator decreases, the 
level of the water rises. The amount of oil remaining in the 
lubricator is indicated by the gage glass 13. Since the lubricator 
is connected to the steam line at two points, the steam pressure at 
these two points balance each other, and the pressure which 
forces the oil out comes from the weight of the column of water 
in the tube 6 and the pipe 3. The length of the pipe 3 should be 
at least 18 inches so as to provide sufficient condensing surface 
and a column of water high enough to force the oil out of the 
lubricator. 

When the level of the water in the lubricator reaches a point 
near the top of the gage glass 13, the lubricator should be refilled 
with oil. This is done by first closing the valve 5 and the valve on 
the feed pipe 11 in order to cut the steam pressure off of the 
lubricator at both top and bottom. The filling plug 15 of the 
lubricator may then be opened, but it must never be opened while 
there is steam pressure on the lubricator, as the hot oil will then 
be blown out and may cause serious injury to persons standing 
near; for this reason every precaution must be taken to close 
valve 5 and the valve on the feed pipe 11, before the filling plug 
15 is removed. The drain cock 14 may now be opened and 
enough water drained out of the lubrictaor until its level is near 
the bottom of the gage glass 13. The oil may then be poured in 
through the filling hole at the top until the lubricator is full, 
after which the filling plug 15 is replaced. 

In order to start the lubricator operating again it is only 
necessary to open the valve 5 and the valve on the feed line 11, 
and then regulate the flow of oil by means of the regulating 
valve 9. 

The hydrostatic lubricator, described above, is not automatic 
in its action, since it must be started and stopped by hand. It 
is difficult to maintain a constant feed, especially at a slow rate, 
because the feed is affected by the changes in viscosity of the oil 
due to changes in temperature. This is particularly true just 
after the lubricator has been refilled. 

These disadvantages are overcome in the mechanically operated 
lubricator, one form of which is illustrated in Fig. 163. This type 
of lubricator is constructed with single and multiple feed delivery 



LUBRICATION 



265 



pipes, the number of feeds being limited to the number of points 
of application demanded in each case. 

In Fig. 163 the oil reservoir is shown at 1, the level of the oil 
being indicated by a glass gage which is not shown in the illus- 
tration. The oil reservoir is provided with a strainer so that fresh 
oil poured into it will be strained and any small particles of solid 
matter which might clog the small oil passages strained out. 
The oil plunger 2, draws oil into the pump on the suction stroke 
and discharges the oil through the nozzle 3 on the delivery 




il i2 



Fig. 163. 



stroke. The oil drops form around a guide wire 4 and rise through 
water in the sight feed. They then pass a nonreturn valve 5, and 
are forced through the check valve 7, at the extreme end of the oil 
pipe 6, into the atomizer 8, located in the steam pipe 10. 

By means of the adjusting nuts 11 and 12, which change the 
stroke of the pump plunger, the oil supply can be varied from the 
smallest amount, say one drop in ten minutes, to practically the 
full capacity of the pump stroke. 

An atomizer, such as shown at 8 in Fig. 163, is the most efficient 
way of spraying oil into and thoroughly mixing it with the steam. 
The atomizer is spoon-shaped on its upper side and has slots 



266 STEAM ENGINES 

extending through it. The steam strikes against the spoon- 
shaped upper surfaces and forces the oil through the slots with 
great velocity, breaking it up into very fine particles which are 
thoroughly mixed with the steam flowing through the pipe. 

The atomized oil is distributed in the form of a uniform coating 
or film over the cylinder walls, valve seats, and piston rod, and 
lubricates these surfaces in a most economical manner. 

Mechanically operated lubricators have several advantages 
over the hydrostatic lubricators. The oil is fed only when the 
engine is running, as the lubricator pump is operated from the 
engine and starts and stops with the engine. This is more 
positive and reliable and a more uniform feed is maintained. 

Certain types of mechanical lubricators, however, are not suit- 
able for compound condensing engines operating on a light load, 
unless the check valves on the discharge side of the pump are 
spring loaded, as a vacuum acting on the oil pipe with atmos- 
pheric pressure on the reservoir will syphon the oil from the 
reservoir. 

Lubrication of Valves. Slide Valve. — As a slide valve is posi- 
tively operated, it can theoretically be operated at any speed; 
but in actual service, due to its unbalanced construction, it is not 
operated at as high speed as the piston valve. 

The flat surface of the slide valve which rubs against the valve 
seat is difficult to lubricate, particularly when the slide valve is 
large. In some extreme cases, oil grooves are cut in the valve or 
in the valve seat to assist in spreading the oil all over the fric- 
tional surfaces. 

The use of the slide valve is restricted to a maximum steam 
pressure of about 120 lb. or a maximum steam temperature of 
about 450° F. This is due to the large, flat, frictional surfaces 
of the slide valve and its seat, and the difficulty of introducing the 
oil thoroughly between them, owing to the great pressure on the 
valve. Excessive steam pressure prevents the formation and 
maintenance of the oil film, resulting in metallic contact of the 
rubbing surfaces, with excessive friction and wear. Excessive 
steam temperature will result in unequal expansion and distor- 
tion of the valve and valve seat. Steam will leak past the valve 
seat, causing cutting of the surfaces. As the oil film is thinned 
out, due to the high temperature, it will be unable to resist the 
pressure between the frictional surfaces, and metallic contact and 
wear will follow. 



LUBRICATION 267 

Upon removing the cover from the steam chest for inspection, 
excessive friction of the sUde valve is always indicated by a dry- 
ness of the rubbing surfaces, which will show wear and bright 
streaks of cutting. Wear on the valve seat will be reduced if the 
cast iron of the valve seat is slightly harder than that of the valve. 

Improper lubrication results in abrasion and cutting; excessive 
leakage of steam takes place and wipes away the lubricating oil 
from the valve seat, making necessary an increased consumption 
of oil. Friction of the slide valve often produces groaning during 
operation, and the excessive resistance in moving the valve causes 
the eccentric rod to vibrate. With efficient lubrication the valve 
operates without noise; the eccentric rod works smoothly; and, 
when inspected, the friction surfaces of the valve and seat have a 
poUshed, dull, glossy appearance. 

Experience has shown that when the oil is introduced into the 
steam line and atomized, it is most thoroughly distributed. In 
many cases, atomizing the oil has furnished the lubrication which 
has overcome groaning and other troubles with slide valves where 
direct methods of application have given insufficient results. 

Corliss Valves. — Although the Corliss valves have a rotating 
motion and the rubbing surfaces are cylindrical, conditions of 
steam pressure and temperature affect their lubrication in the 
same manner as with slide valves. There is this difference, how- 
ever, that insufficient lubrication will cause the sluggish closing 
of the admission valves, since they are not connected directly to 
the operating mechanism during the closing period. Otherwise, 
excessive friction is indicated by groaning and vibration of the 
valve and operating mechanism, as in the case of slide valves. 

Corliss valves should be lubricated by introducing the oil into 
the steam line and atomizing it, as it will then be quickly carried 
to the rubbing surfaces. Introducing the oil directly over the 
valves is wasteful or inefficient or both. 

Piston Valves. — On account of the symmetrical shape of the 
piston valve and sleeve there is but little pressure between them. 
This shape also permits them to expand uniformly under high 
steam temperature. This type of valve is, therefore, well 
adapted to running at very high speeds. The large surface 
moved over by the piston valve at high speeds demands effective 
lubrication, which is best secured by the atomization method. 

Poppet Valves. — Poppet valves have no sliding motion and, 
therefore, do not require lubrication, but the valve stems which 



268 STEAM ENGINES 

move up and down in guides require a small amount of lubricant. 
External lubrication of the valve stems is likely to cause them to 
stick, owing to the fact that the clearance between guide and 
valve stem is very small. For this reason, atomizing the oil and 
using it sparingly will give best results. 

Piston and Cylinders. — In the lubrication of steam-engine 
cylinders we have to consider two types of engines, vertical and 
horizontal. In vertical engines the pressure between the piston 
and cylinder is moderate, being due almost entirely to the spring- 
ing action of the rings. For this reason less oil is required for 
lubrication than with horizontal engines. In horizontal engines, 
the lower part of the cylinder carries the weight of the piston in 
addition to the pressure of the rings. Some large horizontal 
engines have tail rods which, together with the piston rod, carry 
the major portion of the weight of the piston. The duty of the 
piston rings, then, is simply to prevent leakage of steam. The 
pressure of the piston has an important bearing on cylinder 
lubrication because the oil film is at a high temperature and 
excessive pressure may easily destroy it, resulting in friction and 
wear. 

The inside of a properly lubricated steam-engine cylinder will 
have a dull appearance due to the presence of a film of oil. The 
presence of the oil can be detected by wiping the surface with a 
piece of white paper, the stain left by the oil having a brownish 
color. If the stain is black it is an indication that the oil is 
being carbonized. When the film of oil has been wiped away the 
surface underneath should appear dull and glossy. If wear has 
taken place it will be indicated by the surface being bright and 
there will usually be streaks or scratches also. 

Piston and Valve Rods. — Piston and valve rods are always pro- 
vided with stuffing boxes to prevent leakage of steam out of the 
cylinder or valve chest, and, in the case of low-pressure cylinders, 
to prevent the leakage of air into the cylinders when the engine is 
operated condensing. 

Stuffing boxes are packed with either soft or metallic packing. 
Full and efficient lubrication of the packing is essential as a 
perfect seal can only be obtained by the presence of a complete oil 
film on the rods. 

Soft packings are used only under moderate steam conditions. 
The friction between the packing and the rod is always com- 
paratively high, and if the packing is screwed up too tight, it 



LUBRICATION 269 

causes grooving or scoring of the rod, and it is then difficult to 
prevent leakage of steam. 

In reversing engines, the engine is reversed by changing the 
movement of the valves with relation to the position of the pis- 
tons. This is done by hand with small engines and by a special 
reversing engine in the case of large engines. In either case the 
pull required to reverse the engine depends upon the friction of 
the valves moving over their seats and by the friction of the rod 
moving through the stuffing box. Where the valve rods have 
been lubricated externally by the direct method, which is waste- 
ful and inefficient, a change to the atomization method effects a 
marked improvement. The valve rod receives lubrication 
when inside the valve chest and furnishes efficient lubrication to 
the packing. As a result, external lubrication of the valve rods 
can be dispensed with. 

Metallic packing is much superior to soft packing and it should 
be used where the steam temperature is high. The friction of 
metallic packing is less than with soft packing as it exerts only a 
slight pressure, and there is less danger of scoring the piston rod 
with it. With highly superheated steam it is usually necessary 
to supply a small amount of direct lubrication to metallic packing 
in addition to that supplied by the atomization method, but oil 
used in this way should be used sparingly as an excess of oil 
may clog the packing or become carbonized. With only moderate 
superheat, direct lubrication can usually be dispensed with. 

Influence of Operating Conditions. — Stationary steam engines 
usually operate at constant speed, the changes of load being taken 
care of by changing the volume of steam admitted to the cylinder 
or by changing its pressure. Either will reduce the velocity of 
steam flow through the steam main at reduced loads. As this 
will affect the atomization of cylinder oil it should be considered 
in selecting the oil. With full velocity of the steam, heavy oils 
will be readily atomized, but a low velocity of steam requires 
a lighter oil in order to secure thorough atomization and good 
distribution. 

The speed of the engine also affects the lubrication. High- 
speed engines, making short quick strokes, take in only a small 
volume of steam at each stroke, so that only a light-bodied, 
quick-acting oil will give efficient lubrication. 

If the supply of steam to the engine is wet it makes efficient 
lubrication more difficult because the wet steam has a tendency to 



270 STEAM ENGINES 

wash the film of oil off the surfaces of the cylinder. Oil to be used 
with wet steam should have great endurance and must be of such 
quality that it will readily combine with moisture and cling to the 
cylinder walls. 

With compound and triple-expansion engines, even when 
the supply of steam is dry, the drop of pressure causes condensa- 
tion so that the supply of steam to the following cylinders will 
be wet. It is, therefore, necessary sometimes to use one grade 
of cylinder oil for the high-pressure cylinder and a different 
grade for the low-pressure cylinder in order to secure efficient 
lubrication. 

The use of highly superheated steam requires the highest 
quality of oil as the friction and high temperature will carbonize 
and decompose poor quality oils. If the engine operates under 
fairly heavy load a heavier-bodied oil may be used than when the 
engine operates at comparatively light loads. Under the latter 
condition a medium-bodied oil is best. 

Low-grade mineral and compounded cylinder oils are difficult 
to separate from the exhaust steam. The best grades of mineral 
oils separate more easily from exhaust steam and feed water than 
do compounded oils; it will, however, be found that more oil will 
be required to give effective lubrication when using a mineral oil 
than when using a compounded oil. 



CHAPTER XX 
STEAM TURBINES 

General Principles. — The principles of the steam turbine and 
the operation of steam in it are entirely different from the princi- 
ples of the steam engine and the operation of steam in it. In a 
steam engine some of the moving parts have a reciprocating 
motion and the action of the steam is intermittent, while a steam 
turbine is an apparatus in which the moving parts have only a 
rotating motion and the steam which passes through it acts in 
such manner as to produce a constant angular velocity. 

In order to illustrate in a homely way the difference between 
the action of a steam engine and that of a steam turbine, consider 
a large wheel supported in a horizontal plane on a vertical shaft. 
The wheel may be rotated by grasping the rim and walking con- 
tinuously around the shaft. In a similar manner the pressure 
of the steam acts upon the piston of a steam engine and pushes it 
forward. The wheel may also be turned by standing in one spot 
and grasping the rim and moving it with first one hand and then 
the other in a similar manner to opening or closing a large valve 
by hand. In this case the wheel is not turned by exerting a 
pressure at one point on its rim, as in the previous case, but by 
exerting a pressure at first one point on the rim and then another 
so that all points on the rim are used successively. This is the 
manner in which steam acts in a steam turbine. 

In order to make the wheel turn while standing in one spot we 
must have a firm place to stand on and a good grip on the floor in 
order to be able to exert the required pressure on the wheel. This 
means that in every turbine there must be certain stationary 
parts which are firmly fixed to the casing in order that the steam 
may have a good grip on the moving or revolving parts which are 
fastened to the power-transmitting shaft. 

Although we can readily understand how we may turn the 
wheel by standing in one point, it is not so easy to understand 
how steam, which is a flexible medium, can grip the wheel of a 
turbine in such manner as to turn it. 
28 271 



272 



STEAM ENGINES 



There are two methods by which this may be done and these 
will be illustrated by the following examples in which water is 
used instead of steam, remembering that the actions taking place 
with steam will be the same as those taking place with water. 

If a hose nozzle be directed against a board arranged as shown 
in Fig. 164, the water flowing from the nozzle will exert a pressure 
on the board as indicated by the scale at the left. The pressure 
exerted upon the board will depend upon the velocity of the water 
and upon the angle at which the water strikes the board. If the 
stream of water is directed almost parallel with the board it will 
exert only a small pressure but if it is directed so that it strikes 




Fig. 164. 

the board almost perpendicular to its surface it will exert a 
much larger pressure. 

With the arrangement of the nozzle and board shown in Fig. 
164 it will be appreciated that the stream of water will break and 
cause it to splash. This reduces the pressure which the stream 
of water can exert. If, however, the board is curved, as shown 
in Fig. 165, the water will pass over its surface in a smooth stream 
without splashing and at the same time the nozzle maybe directed 
so as to produce the maximum pressure. The pressure exerted 
by the stream of water is due entirely to changing the direction 
in which the stream is flowing. The amount or intensity of the 
pressure will depend upon the velocity of the stream, upon the 
weight of the water (or steam) and upon the angle through 
which it is turned. 



STEAM TURBINES 



273 



The velocity of- the stream when it leaves the board will be 
practically the same as its entering velocit}^, the only loss of 
velocity being that due to the friction of the water passing over 
the surface of the board; this is small if the board is smooth. 
Moreover, if the water leaves the board at the same angle at 
which it strikes, it, that is, if the curvature of the board is 
uniform, the pressure will be directed along the axis of the board 
as indicated by the arrow, and there will be no side thrust. 

The above discussion has shown how a stream of water directed 
along a curved surface can produce a pressure, but in the case 
considered, no work was done because the curved board does not 
move. In order for work to be done, a force must act or move 




Fig. 165. 

through a distance. The product of the force and the distance 
will then be the foot-pounds of work done. If the curved board 
is fastened to the rim of a wheel, as shown in Fig. 165, and the 
wheel is fastened to an axle, then the pressure on the board due to 
the stream of water will cause the wheel to turn and to do work. 
If, instead of only one of the curved boards, or blades as we will 
now call them, a number of blades are placed around the circum- 
ference of the wheel so that as soon as one blade has moved out 
of the stream of water another will move into it, a continual 
pressure will be exerted and the wheel will revolve uniformly. 
Also, instead of only one nozzle and stream of water, there may 
be several nozzles spaced at intervals around the circumference 



274 STEAM ENGINES 

so that instead of the vanes receiving pressure from one nozzle 
they may receive pressure from several nozzles and thus increase 
the total pressure acting upon the rim of the wheel. This con- 
stitutes a turbine similar to one type which is in common use, 
a type known as the impulse turbine. 

In discussing the action of the stream of water on the curved 
vanes it has been assumed that the vanes were standing still. It 
remains now to discuss the effects produced by the movement of 
the vanes, because this is the condition that will exist when the 
turbine is running. This is important because the pressure on 
the vane depends upon the velocity with which the water is 
moving over the surface of the vane and this velocity is affected 
very much by the movement of the vane. In other words, the 



Fig. 166. 

velocity of the stream relative to the vanes is entirely different 
from the absolute velocity of the stream. 

To illustrate this and to show how the velocity of the stream 
relative to the vane, when the vane is moving, may be determined, 
consider the case of a person walking across the floor of a railway 
coach when the coach is moving. Referring to Fig. 166, suppose 
the train is moving forward with a velocity represented by the 
length of the line AB and that the direction of this hne also repre- 
sents the direction in which the point A on the floor of the coach is 
moving. Now suppose a person standing at A walks across the 
coach in the direction AD and walks a distance, measured on the 
floor of the coach, represented by the line AD, in the same length 
of time that the train covers the distance AB. Then the actual 
motion of the person with respect to the ground is represented in 
direction and length by the line AC which is the diagonal of the 
figure A BCD. That is, the line AB represents in length and 



STEAM TURBINES 



275 



direction the absolute velocity of the train with respect to the 
ground, the line AC represents the absolute velocity of the person 
with respect to the ground, and the line AD represents the ve- 
locity of the person relative to the train. 

Applying the same kind of diagram to the case of a stream of 
steam passing over the surface of a vane of a steam turbine we 
would have a figure similar to the one shown in Fig. 167. In this 
diagram the curved line AG represents the curved surface of the 
vane. The line AD represents the velocity of the vane as the 
wheel turns, the line AE represents the velocity of the steam as it 
leaves the nozzle and strikes the vane, and the line AF represents 




Fig. 167. 

the velocity of the steam with respect to the surface of the vane. 
That is, instead of the entire velocity of the steam AE being 
effective in creating pressure against the vane, there is only the 
velocity AF to produce this pressure. 

Consider next what happens to the velocity as the stream leaves 
the vane. This is shown by a similar velocity diagram drawn 
about G, the point at which the steam leaves the vane. As the 
steam passes over the surface of the vane, its direction is changed 
so that if the vane were standing still it would leave in the direc- 
tion GH, tangent to the vane. The hne GH therefore represents 
the velocity of the steam with respect to the vane. The line GH 
will have practically the same length as the line AF because the 
relative velocity of the steam with respect to the vane will be 



276 STEAM ENGINES 

practically the same at the outlet end of the vane as at the inlet, 
the only loss in this velocity being that due to the friction of 
the steam on the surface of the vane. This is small because the 
vanes are made as smooth as possible. The line GJ represents 
by its length and direction the velocity with which the vane is 
moving, and it is, of course, the same as the line AD. The 
diagonal of the diagram, or the line GK therefore represents by its 
length the absolute velocity with which the steam leaves the 
moving vane and its direction shows the direction in which the 
steam leaves the vane. That is, the steam enters the moving 
vane with an absolute velocity of AE and leaves it with an abso- 
lute velocity of GK. 

Having considered the velocities involved in the operation of a 
steam turbine, we are in a position to study the amount of energy 
obtained from these velocities. If we determine the amount of 
energy in the steam entering the vane and subtract from this 
the amount of energy in the steam as it leaves the vane, the differ- 
ence must evidently be the amount of energy given to the vane. 
This represents the energy developed by the turbine wheel, if we 
neglect the small loss of energy that takes place in the vanes. 

If a weight of G pounds of working fluid (steam or water) enters 
the vane per second with a velocity of vi feet per second, its ki- 
netic energy or energy of motion is 

Gv^ 
2g 
in which formula g represents the acceleration due to gravity, 
32.2 ft. per sec. per sec. 

Also the kinetic energy of the working fluid as it leaves the vane 
is 

2g 

The difference between these two quantities is the energy given to 
the wheel, TF, and the loss L, or 

2g 2g "^ ^"^ 

or 

This equation means that we get out of the steam more useful 
energy W, the greater G is, the smaller the losses, and especially 



STEAM TURBINES 277 

the larger Vi^ — V2^ is. With a given inlet velocity the expression 
vi^ — V2^ will be larger, the smaller the absolute outlet velocity V2 
is as compared with the absolute inlet velocity ^i. It is evident 
from Fig. 167 that the absolute outlet velocity V2 will be smallest 
when the hne GK is at right angles to GJ; that is, when the steam 
leaves the vanes at right angles to the direction in which the 
vanes are moving. This, therefore, is one of the conditions for 
maximum efficiency in a steam turbine and the designer chooses 
the angles and velocities so as to secure this result as nearly as 
possible. 

It can be shown mathematically that another condition for 
maximum efficiency is that the circumferential speed of the 
wheel, c, should be approximately one-half of the absolute inlet 
velocity of the steam, vi, or 

It will thus be seen that the inlet velocity is a very important 
quantity. Let us see, therefore, what we may expect this velocity 
to be. 

In water turbines the absolute inlet velocity of a simple impulse 
turbine depends merely upon the height or head of the column of 
water above the level of the turbine wheel. Suppose this head 
to be "h^^ feet, then the theoretical absolute inlet velocity "v^^ is 

V = \^2gh 

This means, for instance, that a particle of water flowing down 
from a height of 100 feet, reaches a velocity of 

V = \/2 X 32.2 X 100 
= 80 feet per second, approximately. 

Taking one pound of water in a height of 100 feet above turbine 
level, we may say that the energy stored up in the pound of water 
is 100 foot-pounds. In falling through the height of 100 feet this 
stored-up energy is changed into energy of motion, so that the 
energy of motion is 

Q^2 I y^ gQ2 

E = -^ = 9 w 00"^ = 100 foot-pounds, approximately. In 

other words, the stored up energy at the top of the column of 
water is equal to the energy of motion at the foot of the column 
(100 foot-pounds) 100 feet below, in accordance with the 
law of constant energy. How is this with a steam turbine. It is 



278 STEAM ENGINES 

exactly the same, but until recently no one thought of comparing 
steam under pressure in a boiler with water under a very high 
head, because such a comparison was not necessary in steam- 
engine practice. 

Suppose we consider one pound of steam in a boiler under a 
pressure of "p^^ pounds per square inch. What is the amount 
of energy stored up in this pound of steam expressed in foot- 
pounds? If we allow this pound of steam to expand to its 
absolute pressure "pi,'^ with a volume of Vi (whereas the volume 
at the pressure "p'^ may be V) then the amount of stored up 
energy L in foot-pounds is 

and during this expansion the volume increases according to the 
law pV"" equals a constant quantity, n being equal to 1.135. 
Under average boiler and condenser conditions this L is 234,000 
foot-pounds. (Steam pressure 150 pounds, vacuum 28 inches.) 
To understand clearly this figure, 234,000 foot-pounds, we 
must bear in mind that the amount of stored up energy in one 
pound of water with a head equal to that of Niagara Falls is 
only 150 foot-pounds. The head of one pound of steam in a 
boiler under 150 pounds pressure against 28 inches of vacuum 

is therefore — -. In. = 1560 times as high as that of Niagara Falls, 

and for such enormous heads steam turbines are designed. 
The equation 

V = y/2gh 

gives us some idea of the velocity this working fluid reaches when 
it flows down this head 

V = \/2X 32.2X234,000 
= 3880 feet per second approximately. 

This means that the absolute inlet velocity of a working fluid 
expanding from 150 pounds per square inch down to 28 inches 
vacuum would be 3880 feet per second. If the circumferential 
velocity of the running wheel is one-half of this it would be 

c = — 2~ ^ 1^^^ ^^^^ P^^ second 

or 116,400 feet per minute 
or 1318 miles per hour 



STEAM TURBINES 



279 



Up to the present time no material, not even the best nickel 
steel, can withstand the enormous strains due to such high cen- 
trifugal force. 

Unless we allow a considerably greater ratio of inlet velocity 
to circumferential velocity than two to one we are unable to 
build a turbine in this way with only one wheel. As soon, how- 
ever, as we increase this ratio, we have to expect an uneconom- 
ically working turbine. Therefore, as it is absolutely necessary to 
build a turbine so it will be as economical as possible, we have to 
seek other ways to reach practical limits of circumferential 
velocity. The best way to do this is shown in the practice of 
water turbines, as for instance, in Massachusetts the manufactur- 
ing industries are using the water power of the Connecticut River 



^£CO/V£> L£V£L 



THIF?D LEVEL 




FOUf^TH LEVet. 



LOW LeVE.L. 



f'l/kn vi esn 

Fig. 168. 



in such a way that they lead the power canal in loops as shown 
in Fig. 168. The circles here represent the water turbines. 
It is plain that the first row of turbines uses the head between high 
level and second level; the second row of turbines uses the head 
between second level and third level, and so on. In short, the 
whole head between high level and low level is split into a num- 
ber of fractional heads between intermediate levels. 

This arrangement of water turbines indicates plainly the way 
to use the enormous heads of steam turbines with practical means, 
and yet follow the fundamental laws as given above. In other 
words, iu order to get an economically working steam turbine 
with moderate circumferential speeds of the running wheel, 
we must use a number of running wheels, all mounted upon one 
shaft, each of these running wheels to be driven by steam flowing 
through a corresponding number of stationary blades where 



280 



STEAM ENGINES 



the steam partly expands and accelerates to a velocity approxi- 
mately twice the circmiiferential velocity. 

One of the first steam turbines that came into extensive use 
in this country makes use of the first principle stated above; 
that is, it has only one running wheel and develops all of its 
power in this single wheel. This turbine is called the De Laval 
turbine. It is made in small and medium sizes up to about 
350 horsepower but it is not well adapted for large units. Hav- 
ing only one running wheel in which the energy of the steam is 
utilized, this wheel must necessarily run at a very high circum- 
ferential velocity (about 1200 feet per second or 72,000 feet per 
minute). 




Fig. 169. 



An illustration of a De Laval turbine wheel and nozzles is 
shown in Fig. 169. This illustration shows the casing of the 
turbine removed so that the construction of the nozzles and 
wheel may be seen more clearly. Instead of only one nozzle, it 
will be seen that four nozzles are used to direct the steam into the 
vanes, as this brings into operation a larger portion of the cir- 
cumference of the wheel and permits four times as much power 
to be developed as if only one nozzle was used. 

The shape of the nozzles is clearly shown in Fig. 169. This 
shape is such that the steam is expanded in the nozzle and its 
pressure energy changed into velocity energy, at the same time 
causing very httle loss of energy in passing through the nozzle. 
When the steam enters the nozzle, it has a high pressure and 
small volume. This is changed in the nozzle so that as the steam 



STEAM TURBINES 



281 



leaves it, it has only a low pressure but occupies a large volume, 
and consequently it leaves the nozzle and enters the vanes with 
a very high velocity. The vanes are spaced uniformly around 
the circumference of the wheel and they are, of course, deep 
enough to accommodate the required volume of steam. 

A study of the diagrammatic sketch shown in Fig. 170 will 
give a good idea of the changes of pressure and velocity that take 
place in a De Laval turbine. The upper part of the sketch 



BOILER 
PRESSURE 



CO/VO^NSER OR 
EX/YAL/ST R=RE5SURE 



STATIONARY 




GO/LER 



pressure: 




k 



CONDE/^SER OR 
~EXHAUST PRESSURE 



Fig. 170. 



shows an elevation of the turbine shaft, the nozzle, and one-half 
of the wheel. Just below this is shown a plan view of the wheel 
and nozzle. At the bottom are shown lines representing the 
steam pressure and velocity in different parts of the turbine, 
the height of these lines above the base Une being proportional 
to the pressure and velocity. It will be observed that the steam 
pressure decreases within the nozzle from boiler pressure to 
exhaust or condenser pressure, that the running wheel revolves 
in this low pressure, having the least possible resistance, and 



282 STEAM ENGINES 

further, that the steam velocity reaches its maximum at the 
outlet of the nozzle and decreases within the vanes of the running 
wheel, thus transmitting the energy of the steam to the turbine 
shaft, and that it leaves the vanes with a certain lost or unused 
velocity. This unused velocity is necessary in any turbine to 
carry the exhaust steam away from the running wheel. 
The above diagram shows further that the pressure is the 
same on both sides of the running wheel, and therefore, the 
axial thrust due to steam pressure is theoretically zero in this 
turbine. 

That class of steam turbines which, instead of utilizing the 
energy of the steam in a single running wheel, divide it among 
several running wheels and thereby permit a much lower cir- 
cumferential velocity, is well represented by the Curtis steam 
turbine, which is used very extensively in this country. These 
turbines are made in all sizes from the smallest to the largest 
and on account of their smaller number of revolutions per min- 
ute they can be connected directly to electrical machinery with- 
out it being necessary to use a set of reduction gears between 
the turbine shaft and the shaft of the driven machinery. 

A diagrammatic sketch of the Curtis turbine arranged similar 
to the one for the De Laval turbine, is shown in Fig. 171. It 
will be observed that the running wheels in this turbine are 
divided into two sets. Each set consists of two running wheels, 
marked R, and a stationary wheel or set of vanes, marked S, 
between them. Steam is directed against the first set of running 
vanes by nozzles similar in shape to those used in the De Laval 
turbine. The stationary vanes between the two running wheels 
are merely for the purpose of changing the direction of the steam 
so that the steam may strike each set of running vanes in the 
same direction. 

In the first nozzles one part of the pressure energy is converted 
into velocity energy. With the velocity corresponding to this 
partial expansion of the steam the first running wheel is impinged 
and one fraction of the whole steam velocity is made available 
in this wheel. Leaving this wheel, the steam is led through 
intermediate or stationary vanes to the next running wheel, 
where the remaining part of the velocity due to the first expansion 
is made available. 

The diagram shows the decrease of the pressure from boiler 
pressure to a medium pressure in the first nozzles, the steam 



STEAM TURBINES 



283 



pressure remaining the same in the whole first part of the turbine 
no further expansion taking place in it. 

After having left the last running wheel of the first part of 
the turbine, the steam is compelled to pass through nozzles in 
which it expands down to condenser or exhaust pressure thereby 
increasing its velocity again up to a certain maximum. This 



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Fig. 171. 

steam jet impinges the first running wheel of the second part 
of the turbine, makes one fraction of its velocity available in it, 
leaves the wheel with a certain velocity, is led with this velocity 
through the following stationary vanes and makes the rest of 
the velocity available in the following running wheel. 



284 



STEAM ENGINES 



The diagram shows the changing steam pressure and steam 
velocity very plainly. Of course, the number of running wheels 
and the number of sets into which they are divided depends upon 
the conditions under which the turbine is to run. In the above 
figure we merely take schematically two sets of wheels with two 
wheels in each set to illustrate the principles of the turbine. 





Fig. 172. 



The intermediate blades in each stage and the stationary 
blade separating one stage from another need openings only 
over a certain part of the circumference. The openings in the 
intermediate vanes increase in radial height according to the 
decreased velocity if there is more than one intermediate row 
in one stage, and the openings of the stationary blades increase 



STEAM TURBINES 



285 



in radial height or in the part covering the circumference ac- 
cording to the increasing volume. 

In the Curtis turbine, as in the De Laval, the steam pressure is 
the same on both sides of each running wheel, hence there is no 
axial thrust along the shaft. 



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Fig. 173. 



The two kinds of turbines described above belong to the general 
type known as impulse turbines. Another well-known type of 
turbine operates by both the impulse of the steam against the 
vanes and by the reaction of the steam as it leaves the vanes. 

An ordinary lawn sprinkler such as that shown in Fig. 172 is a 



286 STEAM ENGINES 

good illustration of the force of reaction exerted by a stream of 
flowing fluid. In this case the streams of water flowing from the 
ends of the arms of the sprinkler exert a backward push or 
*' reaction" which is sufficient to rotate the arms. In the same 
way, steam flowing at high velocity from a properly shaped 
nozzle or vane will produce a reaction. It is this reactive force 
that is used in the type of turbines mentioned above. One of 
the best known of this t3^pe is the Westinghouse-Parsons turbine, 
which is illustrated schematically in Fig. 173. 

No nozzles are used in this turbine, but the rows of stationary 
vanes take the place of nozzles and guide the steam into the 
adjacent rows of running vanes. The outlet opening between the 
running vanes, being smaller than the inlet opening, the steam 
is compelled to expand and accelerate within these running 
vanes exerting thus a back pressure upon them when the steam 
leaves. In this turbine there are therefore two ways in which 
the steam moves one vane. First, the impact at the inlet and, 
second, the reaction at the outlet. On account of the laiger 
angles that must be used in shaping the vanes there cannot 
be made available in each row such a large fraction of energy 
as can be made available by using pure impact and considerably 
smaller angles. 

The steam leaving the first row of running vanes flows through 
the next row of stationary ones, expands there, and impinges 
with the new velocity the second row of running vanes, and so on, 
the vanes increasing in radial height in conformity with the 
expansion hne of the steam and the diameter increasing, too, 
in order to reach higher circumferential velocities. 

The schematical diagram shows plainly that the pressure line is 
nearly one continuous curve from the inlet to the outlet without 
any offsets or stops. The curve of the absolute steam velocity 
explains itself. 

The next consequence of the shape of the pressure line is that 
the running wheels have a higher steam pressure on the inlet end 
than on the outlet end, that is, the whole drum with all the run- 
ning vanes is thrust axially in the direction of the steam flow. To 
balance this thrust, balancing pistons are used, keyseated to the 
turbine shaft and of the same diameters as the different running 
rows, and being connected b}^ steam channels with the space in 
which the corresponding running rows revolve, thus balance 
each running row. 



INDEX 



Figures refer to pages 



Absolute and gage pressures, 60 
Absolute pressure, 60 

vacuum, 60 
Action of slide valve, 9 

of steam engine, 2 
Admission, 3 

line, 96 

pressure, 95 

steam, 125 

valve, 108 
Air pump, 61 
Amount of vacuum, 61 
Angle of advance, 140 
Atmospheric line, 90, 120 

pressure, 59, 60, 64 
Atomizer, 265 
Atomizing, 267 
Automatic high-speed engine, 13 

B 

Back-pressure, 91 
Balanced value, 14 
Barometer, 60 
Bearings, 48 
Block crosshead, 43 
Bore, 30 
Box piston, 37 
Brake constant, 128 

horsepower, 103 

Prony, 104, 126 

rope, 106 
British thermal unit, 57 

C 

Center, 3 

Centrifugal force, 212 
Classification of engines, 6 
Clearance, 4, 180 

volume, 96 
Commercial cut-off, 396, 397 

29 



Commercially dry steam, 70 
Compound expansion. 111 
Compounding, 111, 225, 239 
Compression, 3 
Computations, 126 
Condensate, 119 
Condensation, 111, 198 

of steam, 242 
Condenser, 63, 225 

barometric, 249 

ejector type, 253 

high vacuum, 253 

jet, 247 

purpose of, 240 

siphon, 248 

surface, 250 

Wheeler, 250 
Condensing apparatus, 246 

engine, 7 
Connecting rods, 45 
Consumption, steam, 153 
Corliss engine cylinder, 34 

engines, 17 

valves, 17 
Counterbore, 30 
Crank, 47, 133 

and eccentric, position of, 145 

angle, 141 

circle, 146 

pins, 47 
Crossheads, 42, 87 
Cut-off, 3, 176 
Cycle of steam engine, 2 
Cyhnder, 30 
Cylinder condensation, 107, 225 

measuring, 114 



D 



Dashpot, 199 
Dead center, 3 
Double acting, 4 



287 



288 



INDEX 



Drum, 86 
Dry steam, 66 

equivalent, 121 

line, 115 
Duration of engine test, 122 

E 

Eccentric, 133, 143, 207 

circle, 146 

rod, 143 

strap, 203 

swinging, 173, 177 
Eccentricity, 133, 146 
Effect of heat, 57 
Efficiency, mechanical, 106 

perfect engine, 125 

ratio, 125 

thermal, 126 
Engine, compound, 226, 236 

condensing, 241 

constant, 103 

Corhss, 145, 202, 241 

Cross-compound, 229, 233 

horizontal 4-valve, 122 

Leavitt pumping, 122 

Mcintosh and Seymour, 122 

noncondensing, 241 

plain slide valve. 111, 171 

Rice and Sargent, 122 

Rockwood-Wheelock, 122 

simple, 226 

sUde valve, 222 

tandem-compound, 228 

uniflow, 112 

Westinghouse vertical, 122 
Equal leads, 161 
Events of cycle, 3 
Exhaust, 3, 120 

closure, 91 

lap, 134, 150 

steam, 125 

valve, 91 
Expansion of steam, 93, 227 

line, 96, 120 



Feed water, 119 
Flywheel, 51, 211 
governor, 176 



Formation of steam, 64 

Frame, 27 

Friction, 255 

brake, 103, 118 
horsepower, 106, 130 
load, 118 

G 

Gage, mercury, 62 

pressure, 60, 63, 66 

vacuum, 245 
Girder frame, 28 
Glands, 41 
Governing, 211 

Corliss engines, 20 

high-speed engine, 15 

steam engines, 12 
Governor, centrifugal, 221 

Corliss engine, 208 

Hartnell, 218 

loaded, 218, 220 

pendulum, 212, 215 

Proll, 218 

Rites inertia, 222 

shaft, 20, 221 

throtthng, 182, 213, 217 

Watt, 214 
Gumming test, 260 

H 

Heat of the liquid, 65 

units, 57 
Heavy duty frame, 28 
High pressure cylinder, 212 
Horizontal engine, 6 
Horsepower, 58 

brake, 103, 118 

constant, 127 

indicated, 102 
Hunting, 216 

I 

Impulse turbine, 274 
Indicator, 79, 81, 92 

diagram, 123, 135 

spring, 99 
Inertia governor, 222 
Initial condensation. 111 
Inside admission valve, 141 
Interpolation, 68 



INDEX 



289 



K 

Knock-off lever, 208 



Latent heat, 65, 67, 69 

of evaporation, 65 
Lead of valve, 139 
Links, movable, 85 
Locomotive pistons, 39 
Lubricating external bearings, 261 
Lubrication, 255 

cylinders, 268 

external bearings, 261 

pistons, 268 

steam engine, 261 

valves, 266 
Lubricators, hydrostatic, 263 



M 



Marme engine, 26 

cylinder, 36 

pistons, 39 
Mean effective pressure, 98, 100 
Measare vacuum, 61 
Mechanical efficiency, 106 

equivalent of heat, 58 
Mercury column, 243 
Metallic packing, 41 

N 

Noncondensing engine, 100 
Nonreleasing Corliss engine, 23 
Nozzles, 280 

Numerical relation between heat and 
work, 58 

O 

Oils, testing, 260 

Open rod construction link motion, 

189 
Over-governing, 216 



Pantograph, reducing motion, 89 

Partial vacuum, 60 

Parts of steam engine, 5, 27 



Perfect engine, efficiency of, 125 
Perfect vacuum, 60 
Piston, 37 

dashpot, 199 

displacement, 94, 116 

locomotive, 39 

marine, 39 

position, 144 

rings, 38 

valve, 14 
Planimeter, 98 
Port opening, 148 
Power, 58 
Pressure, 59 

absolute, 60 

atmospheric, 59 

back, 240, 241 

exhaust, 123 

gage, 60, 90 

maximum admission, 96 

terminal, 153 
Properties of steam, 66 
Pump, compound, 131 

small direct acting, 135 

Q 

Quadruple expansion, 226 

R 

Radius bar, 194 
Ratio of expansion, 94 
Reciprocating parts, 1 
Reducing motions, 87 
Re-evaporation, 109, 111 
Regulation of speed, 11 
Release, 3 
Reversing gears, 183 

Woolf, 197 
Rocker arm, 138, 159 
Rods, piston and valve, 268 



S 



Safety cams, 209 
steam, 69, 71 
Second admission, 235 
Sensible heat of steam, 65 
Setting a slide valve, 161 



290 



INDEX 



Shaft governor, 220 
Shifting eccentric, 171 
Simple engine, 6 
Single acting engine, 4 

eccentric valve gear, 200 
Slide valve, 133, 164 

action, 9, 

engine, 8 
Slipper crosshead, 44 
Specific heat, 58 
Speed regulation, 11 
Stability of governor, 214 
Steam consumption, 119, 122, 153 

engine, efficiency of, 123, 124 

jacket. 111 

lap, 134, 150, 192 

line, 90 

meter, 119, 123 

pressure, 60 

superheated, 262 

tables, 65 

turbine, Curtis, 282 

turbine, operation of, 271 
Stephenson valve gear, 195 

link motion, 184 
Stuffing boxes, 40 
Superheated steam, 70, 111, 121 
Surface condenser, 119, 122 



Temperature, 57 

absolute, 125 
Test, acid, 261 

flask and fire, 260 

gumming, 260 
Total heat, 65 
Tram, 159, 202 
Trick valve, 166 
Triple expansion, 226 
True pressure, 60 
Turbine, DeLaval, 280 

impulse, 285 

Westinghouse-Parsons, 286 

U 



Uniflow engine, 112 
Unit of heat, 57 
of power, 58 
U-tube, 62 



Vacuum, 59, 61, 243 

gage, 63 

measuring, 243 
Valve action, 3 

auxiliary, 180 

balanced, 177 

Ball telescopic, 168 

circle, 151 

CorHss, 17, 198, 202, 267 

diagram, Zeuner, 146 

displacement, 143, 146 

gear, 159, 183, 193 

ideal piston, 170 

inside admission, 142 

lead, 148 

Meyer, 179, 182 

multi-ported, 167 

piston, 169 . 

poppet, 267 

rod, 138 

setting, 157 

slide, 266 

steam, 201 

stem, 156, 158 

straight line, 167 

straightway, 83, 84 

three-way, 84 

throttle, 211 

travel, 133 

trick, 166 

with lap, 136 

without lap, 134 
Vane, 275, 282 

stationary, 286 
Vapor pressure, 245 
Variable loads, 118 
Viscosimeter, Saybolt, 260 

W 

Walschaert valve gear, 184, 191, 195 

Wet steam, 69 

Wing crosshead, 42 

Wire drawing, 167 

Woolf reversing gear, 184, 196 

Wrist plate, 198, 207 

Z 

Zeuner valve diagram, 202 



LIBRARY OF CONGRESS 






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